CENTRIFUGAL 
PUMPING  MACHINERY 


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CENTRIFUGAL 
PUMPING  MACHINERY 


THE  THEORY  AND  PRACTICE  OF  CENTRIFUGAL 
AND  TURBINE  PUMPS 


BY 

CARL   GEORGE   de   LAVAL 

Member  American  Society  Mechanical  Engineers,  Naval  Architects  and  Marine 
Engineers,  and  American  Society  of  Naval  Engineers 


McGRAW-HILL   BOOK   COMPANY 

239  WEST  39TH  STREET,  NEW  YORK 

6  BOUVEKIE  STREET,  LONDON,  KG. 

1912 


COPYRIGHT,  1912, 

BY  THE 

McGRAW-HILL  BOOK  COMPANY 


Stanbopc  jprcss 

F.    H.GILSON   COMPANY 
BOSTON,  U.S.A. 


PREFACE 


A  WRITER  upon  centrifugal  pumps  has  said  that  they  defy  the  mathema- 
tician and  possess  more  tricks  than  a  circus  mule. 

This  book  has  been  prepared  with  the  idea  of  supplying  accurate  and 
definite  information,  which  can  be  used  in  actual  design.  The  data  are 
based  upon  experience  in  design,  construction  and  installation  of  this 
type  of  pumping  machinery.  It  attempts  to  set  forth  the  underlying 
principles,  which,  if  properly  applied  and  used,  will  give  the  designer 
enough  information  to  enable  him  to  calculate  results  with  reasonable 
certainty.  Almost  all  books  on  the  subject  in  the  English  language  are 
silent  upon  the  principles  which  govern  the  practical  designer,  and  give 
only  empirical  formulas  which  would  prove  very  costly  in  actual  practice. 

This  book  confines  itself  to  material  which  has  been  used  successfully 
in  practice.  The  author  has  himself  had  charge  of  almost  all  of  the 
installations  described  and  they  become  of  interest,  therefore,  as  records 
of  fact. 

Necessarily,  the  installations  given  are  from  the  practice  of  Henry  R. 
Worthington,  and  it  has  not  been  deemed  wise  or  necessary  to  introduce 
other  practice  to  illustrate  the  principles  which  the  book  seeks  primarily 
to  set  forth.  This  is  not  merely  because  of  the  large  number  of  centrif- 
ugal and  turbine  pumps  which  this  company  has  put  out,  but  because  it 
seems  best  to  confine  the  work  to  actual  experience  and  to  such  installa- 
tions as  have  been  in  service  sufficiently  long  to  be  beyond  the  experi- 
mental stage. 

No  attempt  has  been  made  to  go  into  the  history  of  the  subject,  or  to 
treat  it  from  an  elementary  standpoint,  as  it  is  assumed  that  the  reader 
is  familiar  with  the  laws  of  hydraulics. 

HARRISON,  NEW  JERSEY 
April,  1912. 


V 

241322 


CONTENTS. 


PAGE 

PREFACE >. v 

PART   I. 

CHAPTER  I. 

GENERAL  REMARKS 1 

Low-  and  High-lift  Pumps  —  Low-lift  Pumps  —  High-lift  Pumps. 

CHAPTER  II. 
DIFFUSORS 5 

CHAPTER  III. 
PALANCING  THRUST 7 

CHAPTER  IV. 

PRIMING  AND  FOOT  VALVES 8 

Priming  —  Foot  Valves. 

CHAPTER  V. 
EFFICIENCY 14 

CHAPTER  VI. 

CHARACTERISTICS 20 

CHAPTER  VII. 

OPERATING 27 

Lubrication  of  Bearings  —  Priming  —  Starting  —  Stuffing  Boxes  —  Suction 
Foot  Valve  and  Strainer. 

PART  II. 

CHAPTER  VIII. 

GENERAL  REMARKS 31 

Discussion  of  the  Centrifugal  Pump. 

CHAPTER  IX. 

FIRST  THEORY  OR  ANALYSIS 32 

Theory  of  Impellers  —  Theory  of  Diffusion  Guides  and  Vanes  —  Application 
of  Theory  —  Blades  —  Capacity  of  Centrifugal  Pumps. 

CHAPTER  X. 

SECOND  ANALYSIS  OR  THEORY 48 

Theory  of  Impellers  —  Application  of  Analysis  to  Problem  —  Short  Method 
of  Finding  Characteristics. 

CHAPTER  XI. 

GRAPHICAL  ILLUSTRATION  FOR  DETERMINING  THE  IMPORTANT  ANGLES 60 

Method  of  Correcting  Impeller-vane  Angle. 

vii 


viii  CONTENTS 


CHAPTER  XII. 

THIRD  ANALYSIS  OR  THEORY 66 

Theory  of  Impellers  —  Application  of  Analysis  to  Problem. 

CHAPTER  XIII. 
SCREW  OR  PROPELLER  PUMPS 73 


PART  III. 

CHAPTER  XIV. 

GENERAL  REMARKS 81 

Waterworks    Installation  —  Waterworks  —  Tests    of    Centrifugal    Pumping 
Engines  at  Montreal. 

CHAPTER  XV. 

IRRIGATION  —  DRAINAGE  AND  SEWAGE 89 

CHAPTER  XVI. 
HYDRAULIC  MINING  AND  DREDGING 100 

CHAPTER  XVII. 
MINING  WORK 104 

CHAPTER  XVIII. 

POWER-STATION   WORK 107 

Boiler  Feeding  —  Circulation    of    Water  —  Hot-well    Pumps  —  Centrifugal 
Jet  Condensers. 

CHAPTER  XIX. 

DOCKS 116 

New  Dry-dock  at  Norfolk  Navy  Yard. 

CHAPTER  XX. 
CENTRAL  FIRE-STATION  SERVICE 129 

CHAPTER  XXI. 

FlREBOATS  AND  SHIPBOARD    SERVICE 142 

Fireboats  —  On  Shipboard. 

CHAPTER  XXII. 

SPECIAL  HIGH-SPEED  INSTALLATIONS 147 

CHAPTER  XXIII. 

COMMERCIAL  PUMPS  FOR  GENERAL  INDUSTRIAL  USES 153 


PART   IV. 

CHAPTER  XXIV. 
ELECTRIC  MOTORS 161 

CHAPTER  XXV. 
STEAM  ENGINES  AND  MISCELLANEOUS 164 

CHAPTER   XXVI. 
STEAM  TURBINES 166 

APPENDIX  . .  169 


CENTRIFUGAL  PUMPING  MACHINERY. 

PAKT  I. 


CHAPTER  I. 
GENERAL  REMARKS. 

LOW-  AND  HIGH-LIFT  PUMPS; 

THE  methods  used  for  figuring  and  designing  centrifugal  pumps  are 
usually  regarded  as  more  or  less  mysterious.  This  is  due  to  the  insufficiency 
of  the  data  and  information  available  in  the  English  technical  literature  at 
the  present  time.  Foreign  writers  have  treated  the  theoretical  side  very 
carefully  and  thoroughly.  Private  investigations  have  been  developing 
and  enlarging  upon  the  original  theories  and  showing  how  they  work  out 
in  actual  practice,  but  only  a  few  of  these  results  have  been  placed  in 
the  hands  of  the  public. 

The  centrifugal  pump  presents  many  interesting  phases  which  do  not 
appear  in  any  other  style  of  pumping  machinery,  and  these  must  be  under- 
stood and  their  importance  appreciated  for  the  intelligent  design,  operation, 
or  application  of  the  pump.  Many  of  the  peculiar  features  have  been 
demonstrated  graphically,  and  should  be  carefully  studied  in  Chapter  V  on 
Efficiency  in  this  part.  Some  of  these  peculiarities  afford  a  convenient 
classification  of  centrifugal  pumps  by  characteristics.  The  present  dis- 
cussion, however,  will  classify  the  pumps  as  low-lift  and  high-lift,  accord- 
ing to  the  head  pumped  against. 

LOW-LIFT  PUMPS. 

Although  an  increase  in  speed  conditions  may  change  a  low-head  pump 
to  high-head,  pumps  intended  for  the  latter  service  are  differently  con- 
structed, as  will  be  seen.  Low-lift  pumps  have,  until  recently,  been  little 
understood.  In  design  and  construction  they  have  been  crude  and 
uneconomical,  but  pumps  of  this  type  are  now  built  which  secure  high 
economy  at  both  high  and  low  heads.  This  has  been  the  means  of  putting 
them  into  general  use  for  a  great  variety  of  purposes. 

This  style  of  pump  is  generally  of  the  volute  type.  Under  proper  con- 
ditions it  is  probably  the  lightest  and  cheapest  pump  that  can  be  used  for 
moderate  and  large  quantities  of  water.  The  best  conditions  are  total 

1 


2  :CjENT&lfrtWAL  PUMPING  MACHINERY 

heads  of  from  0  to  150  feet,  short  and  direct  suction  and  delivery  pipes, 
moderate  and  large  quantities  of  water  — never  small  quantities. 

The  chief  applications  for  low-lift  pumps  are :  drainage,  reclamation,  and 
irrigation  work;  waterworks  where  the  lift  is  not  high  and  where  low  first 
cost  is  desirable;  pumping  into  filter  beds;  sewerage  work,  dock  work, 
sluicing,  leakage  in  tunnels,  circulating  water  in  condensers  for  power 
stations,  and  general  water-service  pumps  for  buildings,  mills,  and  heating 
plants.  The  method  of  applying  the  centrifugal  pump  to  these  services, 
and  the  conditions  demanded  for  each  service,  are  explained  in  detail  in 
Part  4,  and  should  be  carefully  studied  by  the  designer  and  engineer  when 
considering  pumps  for  such  installations. 

Centrifugal  pumps  will  deal  with  very  large  volumes  of  water.  Several 
have  been  installed  which  handle  as  much  as  130,000  gallons  per  minute. 
These  pumps  can  be  built  with  a  discharge  pipe  as  large  as  72  inches,  and 
with  pipe  velocities  of  8,  10,  12,  14,  and  16  feet  per  second,  corresponding 
to  the  volume  of  water.  In  fact,  the  amount  of  water  that  can  be  moved 
is  almost  unlimited,  as  there  is  no  difficulty  in  constructing  pumps  for 
these  large  amounts  of  water  for  heads  from  0  to  40  feet  and  with  smaller 
quantities  up  to  150  feet. 

The  ordinary  form  of  low-lift  volute  centrifugal  pump  is  not  adapted  to 
high  speeds,  and  for  these  the  multirotor  pump  or  some  other  type  must  be 
adopted.  These  types  are  described  in  the  chapter,  "Special  High-speed 
Turbine  Installations,"  at  the  end  of  Part  3.  For  very  low  heads  or 
suction  heads  only  and  pumps  working  on  sealed  pipes,  a  combination 
screw  and  centrifugal  pump  or  a  double-screw  pump  is  employed.  This  is 
treated  separately  in  Part  4. 

HIGH-LIFT  PUMPS. 

For  heads  of  more  than  150  feet,  instead  of  the  volute  casing  a  round 
casing  is  used,  which  resembles  that  of  a  water  wheel.  Because  of  this 
resemblance,  high-lift  pumps  are  generally  known  as  turbine  pumps. 

It  has  only  recently  become  known  that  the  simpler  form  of  centrifugal 
pumps,  with  few  parts  and  only  one  moving  piece,  could  be  used  for  fairly 
high  heads.  In  the  Transactions  of  the  Institution  of  Civil  Engineers, 
London,  Volume  XXXII,  it  is  stated  that  an  18-inch  pump  will  work  well 
on  a  20-foot  lift,  and  a  36-inch  pump  on  a  30-foot  lift.  Other  writers  have 
stated  that  the  ordinary  centrifugal  pump  has  a  low  efficiency  at  high  heads. 

The  relation  existing  between  the  actual  pressure  in  the  pump  discharge 
and  the  theoretical  pressure  was  not  known,  and  but  little  attempt  had 
been  made  to  ascertain  it.  In  the  design  of  centrifugal  pumps  this  ratio, 
called  by  some  authorities  the  manometric  coefficient  based  on  experiments, 
must  be  the  basis  from  which  to  work,  together  with  the  percentage  of  the 
useful  work  to  that  expended  in  operating  the  pump. 


GENERAL   REMARKS 


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V 


4  CENTRIFUGAL  PUMPING  MACHINERY 

Practical  results  have  shown  that  single  impellers  without  diffusion  or 
guide  vanes  can  be  made  for  heads  as  high  as  350  feet,  and  it  is  not  unusual 
to  find  commercial  pumps  for  heads  of  150  feet  in  which  the  rapidly  moving 
water  leaves  the  impeller  with  very  little  loss  from  shock  or  eddies.  It  is 
of  course  necessary  to  have  high  peripheral  velocities  in  order  to  obtain 
such  heads.  Figs.  1  and  2  give  results  for  single  impellers  without  diffusion 
vanes  for  heads  from  40  to  135  feet,  in  the  same  pump,  under  different 
speeds. 

Very  high  heads  can  be  obtained  by  multistage  pumps,  which  consist  of 
a  succession  of  rotors  connected  in  series  through  the  casing,  so  that  each 
stage  draws  its  suction  from  the  discharge  of  the  preceding  one  and  so 
raises  the  pressure  progressively.  Pumps  of  the  multistage  type  are  made 
for  heads  as  high  as  200  feet  per  stage  when  proper  conditions  are  met  as  to 
capacity  and  speed.  In  good  practice  the  head  per  stage  varies  between 
100  to  150  feet,  in  order  to  keep  down  the  velocity  of  the  water  so  that  it 
will  not  cause  trouble  from  pitting  the  vanes  or  producing  excessive  wear 
on  the  impeller  and  diffusion  tips.  The  velocity  of  flow  in  the  suction 
inlets  of  the  pump  must  also  be  kept  low,  or  the  water  may  separate  from 
the  entrained  air  and  the  pump  become  noisy,  due  to  the  high  speed  of  the 
impellers  and  imperfect  filling. 


CHAPTER  II. 

DIFFUSORS. 

IN  the  design  and  operation  of  centrifugal  pumps  it  is  important  that  no 
head  should  be  lost  by  shocks  or  abrupt  changes  of  velocity.  The  ordinary 
method  of  taking  care  of  abrupt  changes  in  velocity  consists  in  introducing 
a  whirlpool  chamber  and  volute,  stationary  or  movable  diffusors,  or  stream 
lines.  In  low-lift  pumps  the  casing  is  usually  in  the  form  of  a  volute,  with 
a  gradually  increasing  cross  section.  In  high-lift  multistage  pumps,  each 
section  or  stage  is  fitted  with  a  whirlpool  chamber  or  a  device  with  divided 
stream  passages  known  as  diffusion  vanes  which  divert  the  water  on  leav- 
ing the  wheel  from  a  tangential  direction  into  the  proper  one  for  discharge. 
This  whirlpool  chamber  is  usually  an  Archimedian  spiral,  to  allow  the 
water  to  move  freely  and  to  convert  its  velocity  into  pressure  with  as  little 
friction  as  possible. 

It  has  been  shown  by  experiment  that  a  pump  with  a  proper  whirlpool 
chamber  will  work  with  greater  efficiency  and  against  a  greater  head  than 
one  without  it.  This  is  easily  understood  when  we  consider  that  the  mass 
of  water  revolving  outside  of  the  wheel  has  some  centrifugal  force,  which 
can  be  added  to  that  produced  by  the  wheel  to  increase  the  pumping  head. 
Or,  instead  of  this  chamber,  a  stationary  guide-vane  chamber,  called  a 
diffusor,  can  be  made.  Both  of  these  types  are  intended  to  utilize  the 
energy  of  the  rotating  water  as  it  leaves  the  wheel,  and  increase  the  pump- 
ing power.  The  guide  vanes  or  diffusors  must  be  so  designed  that  the 
stream  lines  will  not  produce  too  much  skin  friction  due  to  additional 
surfaces. 

Efforts  have  been  made  to  have  these  guide  vanes  adjustable  like  those 
of  the  water  turbine,  but  this  has  not  proved  practical.  For  the  best  effi- 
ciency, therefore,  they  should  have  the  shape  suited  to  the  water  path 
corresponding  to  the  conditions  for  which  they  were  designed.  The  vanes, 
in  some  cases,  are  made  as  a  separate  or  removable  casting,  in  others  as 
spiral  grooves  cast  in  the  main  casing,  or  as  passages  in  delivery  compart- 
ments of  the  casings.  The  spaces  in  the  vanes  are  supposed  to  have  the 
shape  which  the  water  would  assume  in  its  passage  from  the  wheel  under 
certain  and  fixed  conditions,  but  the  exact  conditions  of  flow  are  not  known, 
and  as  the  guides  are  always  less  in  number  than  the  vanes  in  the  wheel, 
the  water  is  liable  to  strike  them  at  the  entrances  of  the  ring  and  thus  cause 
injurious  eddies.  Diffusion  rings  can  be  made  without  vanes  and  good 
efficiencies  obtained  even  for  high  heads. 

5 


6  CENTRIFUGAL  PUMPING  MACHINERY 

Investigation  of  movable  diffusors  for  high  heads,  made  to  rotate  freely 
on  the  impeller,  has  also  been  made,  the  results  of  which  show  that  higher 
efficiency  and  heads  can  be  obtained.  This  diffusor  was  originated  by 
Mr.  Barber,  March  31,  1896,  and  by  Professor  Novak  of  Austria  in  1908. 
Its  purpose  is  to  reduce  the  friction  on  the  sides  of  the  runners,  and  to  pro- 
vide ample  whirlpool  for  the  diffusor  chamber  in  the  pump,  doing  away 
with  the  vanes  usually  employed. 

In  the  usual  construction,  the  water  between  the  runner  or  impeller  and 
the  stationary  chamber  has  a  tendency  to  churn,  causing  a  waste  of  energy 
which  increases  with  the  clearance.  A  properly  designed  movable  diffusor 
will  greatly  reduce  this  loss  and  thereby  increase  the  efficiency. 


CHAPTER  III. 

BALANCING  THRUST. 

MORE  or  less  axial  thrust  is  present  in  all  centrifugal  pumps.  It  is 
caused  by  the  unbalanced  pressure  between  the  impeller  and  casing,  be- 
tween the  impeller  and  channel  or  filling-in  rings,  and  also  where  pressure 
acts  upon  unequal  surfaces.  The  water  in  passing  through  the  wheel 
alters  its  velocity  and  pressure,  and  in  changing  from  an  axial  to  a  radial 
direction  produces  a  centrifugal  force  causing  a  thrust.  The  vane  angles 
have  some  influence  upon  the  amount  of  thrust,  dependent  upon  the  veloci- 
ties at  the  inner  and  outer  angles  of  the  vane  with  respect  to  the  impeller. 
The  shape  or  form  of  the  vanes  may  also  cause  a  slight  thrust  under  certain 
conditions.  Very  little  is  known  definitely  of  this  matter,  but  experiments 
have  shown  that  if  we  assume  an  impeller  with  a  cross  section  S,  and  a 

y  2 
velocity  Vz  at  the  hub,  the  axial  thrust  will  be  approximately  3500  S  —  =  T, 

where  T  is  expressed  in  pounds  of  thrust  for  each  impeller. 

Various  arrangements  of  impellers  opposed  to  each  other  have  been  tried 
in  order  to  eliminate  this  axial  thrust.  Impellers  having  two  faces  exposed, 
the  smaller  side  to  the  higher  pressure,  the  larger  one  to  the  lower  pressure, 
have  been  used,  and  by  so  doing  the  thrust  has  been  materially  reduced; 
but,  owing  to  complications  introduced,  this  method  is  open  to  more  or 
less  criticism.  Arrangements  with  bushing  rings,  and  impellers  with 
balancing  holes,  will  materially  aid  in  securing  equal  pressure  upon  the 
areas  within  the  bushing  rings.  For  the  remaining  unbalanced  thrust  a 
marine-type  ball  or  roller  bearing  can  be  used  on  the  outer  end  of  shaft. 
Hydraulic  thrusts  of  different  types  have  been  designed  with  a  view  of 
automatically  adjusting  themselves  to  the  conditions,  and  of  taking  care 
of  the  additional  thrust  that  may  be  produced  by  future  wear  at  various 
points  where  leakage  occurs.  The  most  successful  of  this  type,  with  a 
revolving  steel  disk  having  very  close-running  surfaces,  is  known  as  hydrau- 
lic step  bearing,  and  can  be  operated  by  pressure  from  the  discharge  with 
very  small  loss.  Another  type  consists  of  an  internal  disk  step  located 
either  on  the  discharge  side  of  last  impeller  or  suction  side  of  the  first,  and 
is  designed  to  control  the  thrust  automatically.  An  adjustable  ring  or 
collar  within  the  pump,  securable  so  as  to  control  the  opposing  pressures  in 
the  chambers  or  casings  and  produce  an  opposite  thrust,  has  also  been 
designed. 


CHAPTER  IV. 
PRIMING  AND  FOOT  VALVES. 

PRIMING. 

BEFORE  a  pump  is  started  all  air  must  be  expelled,  and  it  must  be  filled 
completely  with  water.  A  centrifugal  pump  running  in  air  cannot  create 
the  vacuum  necessary  to  raise  the  water  up  to  the  impeller.  Some  pumps, 
particularly  those  handling  hot  water  or  any  liquid  giving  off  a  vapor, 


Steam 


Ejector 


Steam 


"Valve 


Steam 


Fig.  2A.     Methods  of  Priming. 

should  always  be  so  placed  that  the  water  will  flow  to  the  pump;  otherwise 
the  vapor  will  collect  in  the  pump  and  it  will  cease  pumping.  Pumps  so 
located  that  water  will  not  flow  to  them  must  be  primed.  Methods  of 
priming  are  illustrated  in  Fig.  2A. 

There  are  two  principal  methods  of  priming,  i.e.,  by  producing  a  vacuum 
with  an  air  pump,  or  by  pumping  water  into  the  casing;  and  all  the  devices 

8 


PRIMING  AND  FOOT  VALVES 


9 


used  are  different  forms  of  one  of  these  methods.  The  conditions  will 
determine  whether  steam,  compressed  air,  or  water  should  be  used.  A 
simple  method  is  to  have  a  foot  valve  on  the  end  of  the  suction  pipe,  and 
to  fill  the  casing  through  a  small  hole  at  the  top  with  city  water  or  other 
water  under  pressure.  A  second  method  uses  the  siphon  principle.  A 
foot  valve  is  placed  at  the  end  of  the  suction  pipe,  a  gate  valve  in  the  dis- 
charge close  to  the  outlet,  where  it  extends  horizontally  from  bottom  of 
pump,  an  air  cock  at  the  top  of  the  casing,  and  water  is  supplied  from  a 
connection  on  the  side.  When  water  appears  at  the  top  pet  cock  the 
pump  is  primed.  After  closing  the  water  supply,  open  the  discharge 
gradually.  This  initial  priming  is  sufficient  for  all  future  starting,  pro- 


Fig.  2B.     Priming  Device  for  Small  Centrifugal  Pumps. 

vided  that  the  foot  valve  is  tight,  and  that  the  discharge  gate  valve  is 
gradually  closed  as  the  pump  is  shut  down.  The  water  will  then  remain 
in  the  pump. 

For  heads  of  over  30  feet,  it  is  usual  to  place  a  check  valve  in  the  dis- 
charge in  order  to  prevent  shock  on  the  pump,  and  it  is  customary  to  place 
a  foot  valve  on  the  suction  and  a  by-pass  for  priming  between  the  dis- 
charge, above  the  check  valve,  and  the  suction  pipe  above  the  foot  valve. 

Another  method  for  use  with  steam  or  compressed  air  is  to  pump  water 
into  the  casing  by  an  injector.  This  should  be  so  placed  that  it  is  within 
easy  suction  lift  of  the  water.  Priming  without  a  foot  valve,  such  as  is 
necessary  with  pumps  operating  on  wells  where  it  is  impossible  to  place 
one,  requires  a  check  valve  in  the  suction  pipe  close  to  the  pump  opening. 


10 


CENTRIFUGAL  PUMPING  MACHINERY 


An  ejector  is  connected  at  the  top  of  the  casing  and  its  suction  pipe  tapped 
into  the  main  suction  under  the  check  valve. 

A  method  of  priming  with  a  foot  valve,  and  a  check  valve  in  the  discharge, 
using  an  ejector  as  an  exhauster,  is  also  employed.  For  this  method  both 
steam  and  compressed  air  are  used. 

On  large  engine-driven  centrifugal  pumps,  running  condensing,  the  top 
of  the  casing  can  be  connected  with  the  condenser  and  a  sufficient  vacuum 

created  for  priming.  In  such  installations  it  is 
best  to  fit  a  glass  water  gauge  on  the  top  of  the 
pump  so  that  the  operator  may  know  when  the 
casing  is  full  and  prevent  the  water  from  going  into 
air  pump,  particularly  if  this  is  of  the  dry,  rotative 
type. 

For  small  centrifugal  pumps,  up  to  and  includ- 
ing 12  inches,  a  specially  designed  priming  device 
may  be  used.  (See  Fig.  2s.)  It  consists  of  a 
fitting  or  casting  placed  directly  against  the  suc- 
tion opening  and  takes  the  place  of  the  usual 
elbow.  It  is  fitted  with  a  small  hand  pump,  and 
a  clapper  or  foot  valve  in  the  main  passageway. 
Water  is  drawn  through  the  main  opening  into 
the  hand  pump,  and  forced  out  through  a  small 
check  valve  into  the  main  pump,  the  water  being 
retained  there  by  the  main  check  valve  in  the 
primer.  In  order  to  prevent  losses  from  friction 
in  this  type  of  primer,  the  openings  are 
made  very  large. 

A  similar  type  of  primer  providing 
the  same  features  consists  of  an  ordi- 
nary hand  pump  attached  to  the 
suction  opening  and  connected  with 
the  water  supply.  A  foot  valve  is 
required  on  main  suction. 

Still  another  method  is  to  attach  a 
hand  air  pump  onto  the  pump  at  any 

place  below  the  discharge  gate  valve  for  exhausting  the  air  between  foot 
valve  and  discharge  valve. 

In  larger  installations,  separately  driven  electric  or  steam-driven  vacuum 
pumps,  operated  automatically,  are  used,  particularly  where  the  suction 
pipes  are  large  and  quite  long. 

In  mining  installations  it  is  necessary  with  sinking  pumps  to  use  an 
automatic  repriming  arrangement,  illustrated  in  Figs.  3  and  4,  consisting 
of  a  foot  valve  in  the  suction  and  an  automatically  operated  check  valve 


Fig.  3.     Automatic  Priming  Arrange- 
ment for  Sinking  Pumps. 


PRIMING  AND  FOOT   VALVES 


11 


Section  in  Balanced 
1       By-Pass  Valve  B 


Fig.  4.    Automatic  Priming  Arrangement  for  Sinking  Pumps. 


12  CENTRIFUGAL  PUMPING  MACHINERY 

in  the  discharge,  with  necessary  air-relief  valves  at  the  highest  point  of  the 
entrance  pipe  to  the  pump.  Its  purpose  is  to  discharge  the  air  and  to 
refill  the  suction  pipe.  The  illustration  shows  the  automatic  check  valve, 
which  is  wide  open  when  pumping,  allowing  a  free  discharge,  at  the  same 
time  closing  off  the  connection  between  the  delivery  pipe  and  the  entrance 
to  the  first  impeller.  Should  the  pump  stop  for  any  reason,  the  check 
valve  closes  the  connection  from  the  last  impeller  to  the  column  and  allows 
the  water  in  the  column  pipe  to  flow  through  the  small  pipe  and  submerge 
the  impellers,  priming  the  entire  pump  down  to  the  foot  valve.  On 
starting  up  any  accumulation  of  air  in  the  suction  pipe  will  be  discharged 
through  the  automatic  air-relief  valve.  When  all  the  air  has  been  dis- 
charged the  air-relief  valve  closes. 

It  is  advisable  in  this  design  of  pump  to  introduce  the  water  on  top 
rather  than  on  the  bottom.  The  location  of  the  automatic  relief  valve 
allows  the  pump  to  free  itself  from  air  before  the  water  enters  the  first 
impeller.  The  automatic  operation  of  the  discharge  check  valve  can  be 
accomplished  in  various  ways,  as  shown  in  the  illustrations. 

FOOT  VALVES. 

So  much  trouble  has  arisen  in  practice  because  engineers  have  not 
appreciated  the  importance  of  a  properly  designed  foot  valve  for  a  centrif- 
ugal pump  that  it  has  been  deemed  necessary  to  devote  a  separate  section 
to  this  subject. 

The  construction  of  a  centrifugal  or  volute  pump  is  weak  in  itself,  as  the 
pumping  head  is  formed  at  the.  periphery  of  impeller,  and  when  the  pump 
is  working,  the  pressure  on  the  side  plates  is  much  lower.  These  cannot 
be  stayed  in  this  form  of  pump,  and  are  not  supposed  to  stand  the  internal 
strain  or  pressure  due  to  the  total  pump  head.  A  sudden  stoppage  of  the 
column  of  water  traveling  through  the  pipes  causes  a  heavy  pressure  on 
the  sides,  tending  to  open  and  rupture  the  casing.  The  column  of  water  is 
supported  by  the  rotation  of  the  impeller,  and  if  from  any  cause  this  rota- 
tion suddenly  ceases,  the  intensity  of  the  reaction  or  shock  is  dependent  on 
the  weight  of  water  in  the  pipes,  and  on  the  velocity  acquired  by  the  re- 
turning column  before  it  is  finally  arrested.  The  sudden  closing  of  the 
foot  valve  is  frequently  sufficient  to  split  the  pump  casing,  pipe  and  heads. 
This  danger  can  be  reduced  by  furnishing  the  foot  valves  with  a  relief  or 
safety  valve.  This  is  particularly  necessary  when  priming  water  is  used 
under  a  pressure  heavier  than  that  for  which  the  pump  was  built. 

A  relief  valve  of  this  kind  can  be  made  a  part  of  the  foot  valve,  or  it  can 
be  attached  to  suction  pipe  and  discharge  back  into  the  well.  It  is  not, 
however,  advisable  to  employ  foot  valves  on  large  pumps,  and  other  means 
for  priming  the  pumps  should  be  used. 

Foot  valves  should  have  at  least  150  per  cent  of  the  area  of  the  suction 


PRIMING  AND   FOOT   VALVES 


13 


pipe,  which  should  be  the  next  size  larger  than  discharge.  Thus  a  pump 
with  10-inch  discharge  should  have  12-inch  suction  pipe,  and  if  a  foot  valve 
be  employed  its  area  should  be  one  and  one-half  that  of  the  12-inch  pipe  to 
reduce  the  frictional  resistance  through  the  system;  otherwise  these  losses 


Fig.  5.     Centrifugal  Foot  Valve. 

Dsorb  a  large  percentage  of  the  total  work  in  low-lift  pumps,  which  means 
poor  economy,  preventable  with  proper  sizes  of  piping. 
The  valves  should  be  of  the  flap  design,  so  made  that  when  open  they 
rest  on  the  sides  of  the  body,  thus  allowing  a  clear  passage  for  water  through 
le  center  of  the  valves,  as  illustrated  in  Fig.  5. 


CHAPTER  Y. 
EFFICIENCY. 

THE  word  efficiency  in  connection  with  centrifugal  pumps  has  become 
very  ambiguous,  and  has  led  to  many  disputes  in  connection  with  guaran- 
tees and  contracts.  This  ambiguity  has  been  brought  about  by  the  fact 
that  the  word  has  been  used  without  modification  to  designate  the  efficiency 
of  the  pump  only,  of  the  pump  and  prime  mover,  and  of  the  entire  plant 
measured  back  to  the  boiler.  This  uncertainty  can  and  should  be  elimi- 
nated. The  efficiency  of  a  centrifugal  pump,  when  not  otherwise  modified, 
can  mean  but  one  thing,  —  the  ratio  of  the  water  horse-power  output  at  the 
pump  to  the  brake  horse-power  input  at  the  coupling  or  pulley. 

The  water  horse-power  output  is  determined  by  the  total  head  against 
which  the  water  is  pumped  and  the  quantity  of  water  delivered.  The 
usual  method  of  finding  the  total  head  is  to  place  a  gauge  on  the  suction 
line  close  to  the  pump,  another  on  the  discharge,  and  to  note  the  vertical 
distance  between  the  gauges.  The  algebraic  difference  between  the  gauge 
readings  plus  the  vertical  distance  between  the  gauges,  all  in  feet,  is  con- 
sidered the  total  head.  Another  method  is  to  add  to  the  head  thus  ob- 
tained the  velocity  head  in  the  discharge  pipe.  Still  another  method  is  to 
add  to  the  difference  in  gauge  readings  and  the  vertical  distance  between 
gauges,  the  difference  between  the  velocity  heads  in  the  suction  and  dis- 
charge pipes.  It  is  therefore  important  in  considering  efficiencies  that  the 
total  head  be  clearly  interpreted.  /\^ 

The  three  methods  of  obtaining  the  total  head  may  be  represented  by 
the  following  formulae : 

H  =  Total  head  in  feet. 

HI  =  Discharge  head  in  feet. 

Hi  =  Suction  head  in  feet. 

A   =  Vertical  distance  between  gauges  in-  feet. 

HS  =  Velocity  head  in  discharge  pipe  in  feet. 

H4  =  Velocity  head  in  suction  pipe  in  feet. 
First,     H  =  H1-H2  +  A. 
Second,     H  =  HI  —  H2  +  A  +  H%. 
Third,    H  =  H,  -  H2  +  A  +  (#3  -  H*). 

HZ  and  H^  are  made  up  of  the  flow  in  feet  per  second  through  the 
pipes  thus : 

TT       (velocity  in  feet  per  second)2 

3=  ~W 

in  which  g  is  the  acceleration  due  to  gravity,  or  32.2. 

14 


EFFICIENCY  15 

Third  formula  represents  the  actual  head  pumped  against  and  should 
always  be  used  since  the  gauge  reading  shows  the  difference  between  the 
total  head  above  the  gauge  and  the  velocity  head.  Where  the  suction 
and  discharge  pipes  are  of  the  same  diameter,  the  velocity  heads  are  equal, 
and  equation  (3)  becomes  the  same  as  equation  (1). 

High  efficiency  in  pumps  is  obtained  by  changing  the  kinetic  energy  of 
the  water,  as  it  issues  fro.m  the  wheel,  into  pressure,  by  reducing  water 
friction,  such  as  churning  in  the  chambers,  and  the  skin  friction  of  rotating 
disks,  and  by  reducing  the  friction  of  the  bearings.  The  elements  of  total 
pump  efficiency  are  therefore  the  absolute  hydraulic  efficiency,  the  mechani- 
cal efficiency,  and  the  volumetric  efficiency. 

The  absolute  hydraulic  efficiency  r,h+  expresses  the  ratio  between  the  use- 
ful head  and  the  total  pumping  head.  If  h  represents  the  former -and  H  the 
latter,  then  the  total  head  H  is  made  up  of  h  and  all  frictional  and  shock  losses 
of  water  in  the  pump,  and  is  the  most  important  factor  to  be  considered. 
Information  relating  to  these  losses  is  very  meager  and  incomplete.  The 
most  serious  loss  is  due  to  skin  friction  between  the  impeller  and  the  water. 
The  head  pumped  against  varies  as  the  square  of  the  velocity;  hence  the 
wasted  power  varies  as  the  cube  of  the  head.  As  the  head  increases,  the 
loss  from  skin  friction  increases  at  a  more  rapid  rate  and  thereby  imposes  a 
limit  on  the  head  against  which  the  pump  can  be  economically  operated. 
The  work  wasted  in  disk  friction  varies  as  the  square  of  the  radius,  hence  a 
smaller  impeller  at  a  higher  number  of  revolutions  absorbs  less  power  in 
friction  than  a  larger  one  at  a  less  number  of  revolutions  but  with  the  same 
peripheral  velocity.  The  disk  friction  of  the  water  near  the  axis  is  lower 
than  that  at  the  outer  surface  of  the  impeller.  Experiments  have  been 
made  abroad  on  the  power  lost  by  skin  friction  and  the  following  formula 
has  been  obtained: 

W  =  8132  X  -2  X  h2-5  foot  pounds. 

W  =  Power  due  to  resistance  in  foot  pounds, 
n  =  Revolutions  per  minute, 
h  =  Head  in  feet. 

The  largest  losses  by  surface  friction  are  along  the  walls  of  the  casings, 
and  were  the  surfaces  of  the  stationary  walls  and  impellers  alike  the  water  ^ 
would  receive  a  rotating  motion  one-half  that  of  the  impeller.  These 
losses  are  reduced  by  having  the  internal  walls  smooth  and  impellers 
polished,  and  by  proper  clearance  between  impellers  and  casings.  In 
addition  to  surface  friction  there  are  losses  due  to  molecular  friction  in  the 
water  itself. 

The  mechanical  efficiency  depends  upon  the  friction  between  the  shaft,  \/ 
impeller,  etc.,  and  their  bearings,  and  is  a  function  of  the  workmanship, 
fit,  and  lubrication. 


16  CENTRIFUGAL  PUMPING  MACHINERY 

The  volumetric  efficiency  is  the  ratio  of  the  amount  of  water  discharged 
to  the  amount  entering  the  pump.  The  difference  is  the  loss  due  to 
leakage  in  running  fits.  The  greatest  loss  in  a  turbine  pump  is  between 
the  impeller  and  the  diffusion  ring.  These  losses  vary  from  2  to  10 
per  cent. 

Efficiencies  on  large  pumps  have  been  obtained  as  high  as  90  to  92  per 
cent,  and  in  multistage  turbines  as  high  as  85  to  87  per  cent.  This  has 
been  accomplished  by  careful  analysis,  as  shown  above,  by  reducing  the 
internal  losses  and  skin  friction,  by  dissipating  shocks  and  disturbances 
and  turning  more  of  the  velocity  into  effective  head.  The  advance  which 
this  involves  is  shown  by  the  fact  that  until  recently  the  best  efficiency 
obtained  from  the  ordinary  volute  was  40  per  cent.  The  conditions  and 
approximate  efficiencies  which  may  be  expected  with  correctly  designed 
impellers  are  as  follows:  "C 

Capacities  from  75  gallons  to  250  gallons  will  give  about  55  to  65  per  cent  efficiencies ; 

250  gallons  to  900  gallons,  70  per  cent; 

900  gallons  to  3000  gallons,  70  to  73  per  cent; 

3000  gallons  to  6000  gallons,  73  to  75  per  cent;  and 

6000  gallons  to  10,000  gallons,  75  to  78  per  cent. 
Above  this  we  obtain  from  75  to  85  per  cent  efficiencies. 

A  side  entrance  or  single  suction  pump  will  give  slightly  less  in 
each  case.  Sizes  of  the  discharge  pipes  of  pumps  for  the  above  vary  from 
4  to  60  inches  diameter.  Speeds  for  the  small  pumps  vary  from  as  high  as 
3500  revolutions  per  minute  for  the  smallest  down  to  600  revolutions  for 
8-  to  10-inch  pumps.  The  larger  may  run  as  slow  as  150  revolutions  with 
good  efficiency  under  favorable  conditions. 

The  losses  at  the  impeller  outlets  are  due  to  the  discharge  velocity  and 
can  be  expressed  as  (aU2  2  </),  where  U  is  the  absolute  velocity  of  discharge 
from  the  vanes,  and  a  is  a  coefficient  with  a  value  between  0.5  to  0.6. 

The  skin  friction  of  the  rotating  impeller  varies  as  the  square  of  the  pe- 
ripheral speed,  and  also  as  the  area  of  the  casing.  Considering  first  the  loss 
of  head  on  entering  the  impeller  vanes,  it  must  be  assumed  that  the  water 
enters  the  inner  portion  of  the  wheel  radially.  In  order  to  avoid  shock,  the 
direction  of  the  vane  must  coincide  with  the  resultant  of  the  radial  velocity 
of  the  water  and  the  tangential  velocity  of  the  inner  circumference  of  the 
wheel. 

The  loss  of  head  caused  by  changing  of  velocity  at  entrance  of  vane  is 
C/2  2  g  =  y  -  Vz  cot  B. 

The  loss  of  head  at  outlet  of  impeller  is 

(73)2  2  g  =  V  -  Vs  (cot  a  +  cot  p). 

The  remaining  losses  are  due  to  friction  and  are  proportional  to  the  square 


EFFICIENCY 


17 


of  velocity  of  flow  through  impeller  and  to  the  square  of  peripheral  speed, 
and  can  be  expressed  as  follows: 


K  and  L  being  constants. 

The  discharge  head  will  decrease  as  the  capacity  is  increased,  the  head 
depending  upon  Vi  and  Vz,  and  is  represented  by  the  following  formula: 


where  M  and  N  are  constants. 

Therefore  the  total  actual  head  equals  the  theoretical  heads  less  the 
losses.  The  constants  here  mentioned  should  be  determined  from  actual 
experiments  of  the  particular  type  of  pumps  considered.  A  graphical 
illustration  is  shown  in  Fig.  6. 


Velocity  of  Water 

relative  to  the  Whee 

at  Inlet 


/I  —  Vetocity  Impressed  by  Vane  on  Water 
and  propels  the  Water  in  direction  to 
C  instead  of  to  A. 

B~C   is  Velocity  of  Water  relative  to  Impeller 
*t  Entry  of  Vane. 


E"^  b  the  Velocity  of  the  Water  relative 
to  the  Wheel  at  Exit.  O~C    IB  tlie 
Absolute  Telocity  of  Flow  along 
Diffusion  Vanes. 


Velocity  of  Whirl 

Fig.  6.     Diagram  Showing  Loss  of  Head  by  Change  of  Velocity  at  Entrance  of  Vanes 

and  at  Outlet. 


18 


CENTRIFUGAL  PUMPING  MACHINERY 


paadg 


EFFICIENCY 


19 


.  .     li  ||| 

lililii 

J  lliiifllk 

1     sl||  1 1    SS-.  *v 

I  jjjjll* 

I     vi 


H 


CHAPTER  VI. 
CHARACTERISTICS. 

'EVERY  impeller  has  a  fixed  relation  between  the  head,  capacity,  and 
speed,  known  as  the  impeller  characteristic.  The  prevailing  idea  that  an 
impeller  can  be  used  with  only  one  condition  of  head  and  capacity  for  a 
given  speed  is  erroneous.  An  impeller  is  good  for  a  certain  range,  and,  in 
order  to  determine  this,  experiments  must  be  made  and  curves  plotted. 
These  relations  between  capacity  and  head  with  the  speed  constant,  head 


1    I 


S     I     3     3     3     9 

Capacity  in  G.P.M. 

Fig.  14.     Actual  Curves  from  6-inch  Pump. 


and  speed  with  the  capacity  constant,  and  capacity  and  speed  with  the 
head  constant,  are  of  utmost  importance,  and  should  be  considered  with 
the  efficiency  curves  of  the  impeller,  in  the  determination  of  a  proper  design. 
A  series  of  pumps  of  different  sizes  should  be  tested  with  reference  to  the 
following  items: 

(a)  Inside  and  outside  diameters  of  impellers; 

(b)  Outside  and  inside  angles  of  vanes; 

(c)  Radius  and  number  of  vanes; 

(d)  Width  of  impeller. 

20 


CHARACTERISTICS 


21 


•<TH  ' 


22 


CENTRIFUGAL  PUMPING  MACHINERY 


From  these  tests  can  be  found  that  efficiency  which  will  give  with 
varying  speed  and  constant  capacity,  the  most  advantageous  peripheral 
speed  for  the  given  angles;  and  also  that  efficiency  with  varying  capacity 
and  constant  speed,  which  will  give  the  capacity  most  advantageous  for  a 
given  width  of  impeller  and  inside  and  outside  angles  of  vanes.  / 

Fig.  7  shows  a  set  of  characteristic  curves  for  these  variables  and  con- 
stants with  their  respective  efficiency  curves  in  broken  lines.  Fig.  8 
shows  a  set  of  speed  and  head  curves  with  capacity  constant  for  one  im- 


Fig.   19.     Six-inch  Pumps  from  which  Actual  Performance  of  Curves  for  Figs.  14  to  18 

were  taken. 

peller.  The  dotted  efficiency  curves  correspond  to  the  speed-head  curves 
bearing  the  same  numbers.  Fig.  9  shows  a  set  of  capacity-speed  curves 
for  constant  head.  Fig.  10  shows  a  set  of  capacity-head  curves  for  con- 
stant speed.  Fig.  11  gives  characteristics  for  a  regular  volute  pump  which 
can  be  compared  with  turbine  or  diffusion  pumps. 

Figs.  12  and  13  show  a  series  of  capacity-head  curves  for  constant  speed 
upon  which  have  been  interposed  a  series  of  constant-efficiency  curves. 
The  method  of  plotting  these  efficiency  curves  is  shown  at  the  top  of  Fig. 
12.  From  the  capacity-efficiency  curve,  a  given  efficiency  A  is  projected 
to  the  corresponding  capacity  ordinate  on  the  capacity-head  curve  at  A, 


CHARACTERISTICS 


23 


and  points  of  equal  efficiency  for  the  various  speed  characteristics  are  then 
connected.     A  series  of  these  equal-efficiency  curves  will  show  the  maxi- 


Capacitin  G.P.M. 
Fig.  20.     Actual  Curves  from  a  Special  Designed  Pump. 


7*1 


Fig.  25.     Special  Designed  Pump  from  which  Curves  for  Figs.  20  to  24  were  taken. 


mum  efficiency  for  any  speed  with  a  given  impeller,  and  is  very  useful,  since 
the  usual  condition  is  constant  speed. 

For  constant  speed  the  relations  usually  required  are  those  existing  be- 
tween capacity  and  head,  capacity  and  power,  and  capacity  and  efficiency. 


24 


CENTRIFUGAL  PUMPING  MACHINERY 


•<rira 


CHARACTERISTICS 


25 


Figs.  14  to  18  inclusive  show  curves  taken  from  actual  tests  of  a  6-inch 
volute  type  of  pump  illustrated  in. Fig.  19.  Figs.  20  to  24  show  curves 
taken  from  a  6-inch  special  type  of  pump  illustrated  in  Fig.  25. 

These  illustrations  show  fully  the  characteristics  of  these  pumps  and 
illustrate  the  possibilities  of  a  careful  analysis  of  the  subject,  and  point  out 
the  way  to  reach  results  by  the  simplest  and  most  reliable  road. 
^The  characteristics  show  advantages  of  centrifugal  pumps  not  possessed 
by  any  other  type  of  pump.     It  is  impossible  to  produce  a  pressure  in  the 
pipe    lines   higher   than  'that 
corresponding  to  the  speed  as 
shown  from  the  form  of  the 
curve,  obviating  all  danger  of 
line  breaks  when  a  valve  on  the 
discharge   line  is  accidentally 
closed.      The    capacity-horse- 


1200  R.P.  M. 


Gallons  per  Minute 

Characteristic  Curve  for  Fixed  Overload 
Condition. 


power    curves    show    that    a 

sudden  closing  of  the  discharge 

valve  will  reduce  the  load  on 

the  motor.    On  the  other  hand, 

the  usual  types  of  centrifugal 

pumps    overload   the    motors 

should  a  break  occur  in  pipe  Fig.  26. 

line  reducing  the  pump  head 

and   increasing   the   capacity, 

and  in  order  to  care  for  this  the  impellers  should  always  be  designed  so 

that  not  more  than  25  per  cent  overload  can  be  thrown  on  the  motor. 

It  is  to  be  noted  that  when  the  head  pumped  against  is  reduced  an  in- 
crease in  the  discharge  takes  place,  due  to  the  fact  that  the  surplus  head 

produced  by  the  pump  is 
converted  into  a  velocity 
head.  Increasing  the 
speed  may  also  overload 
the  motor,  but  the  im- 
peller can  be  so  designed 
that  25  per  cent  can  be 
made  the  maximum  over- 


10  20  30  40  50  60  70  80  90  100  110  120  130  140  150 160  170 


Fig.  27. 


Characteristic  Curve  for  Fixed  Overload 
Condition. 


load.    /The  characteris- 
tics of  such  pumps  are 

shown  in  Figs.  26  and  27, 

from  which  it  can  be  seen  that  the  maximum  power  at  constant  speed  for 
greatest  variation  in  head  is  only  1.25  times  the  power  absorbed  by  the 
pump  when  giving  the  highest  efficiency. 
\   A  pump  can  be  so  designed  that  the  power  curve  will  be  nearly  constant 


26  CENTRIFUGAL  PUMPING  MACHINERY 

while  the  efficiency  is  maintained.  The  ordinary  design,  however,  gives  a 
capacity  varying  from  70  to  125  per  cent,  with  a  variation  of  head  of  25 
per  cent,  and  an  efficiency  variation  of  about  5  per  cent.  / 

In  Figs.  55  and  58  the  various  forms  of  impeller  vanes  show  the  different 
characteristics  as  usually  applied  to  commercial  problems.  No.  3  will  main- 
tain with  constant  speed  a  uniform  efficiency  under  a  varying  head.  No.  2 
will  produce  a  constant  head  for  the  best  efficiency,  but  has  the  disadvan- 
tage that  the  power  increases  rapidly.  No.  1  would  overload  considerably 
and  is  limited  to  but  few  conditions,  as  when  the  load  is  variable,  due  to 
the  frictional  resistances  other  than  the  regular  pumping  head.  The  extent 
to  which  a  designer  can  go  in  the  development  of  the  shape  of  the  vanes 
depends  entirely  upon  the  conditions  of  service,  and  should  be  carefully 
investigated. 


CHAPTER  VII. 
OPERATING. 

MUCH  trouble  may  be  experienced  in  operating  centrifugal  pumps  if  due 
consideration  is  not  given  to  the  physical  conditions  of  the  pipe  line  and 
installation  and  to  the  material  to  be  pumped.  The  following  paragraphs 
should  be  carefully  considered  before  erecting  and  starting  a  centrifugal 
installation. 

Suitable  foundations  are  essential  for  satisfactory  and  continuous  opera- 
tion. The  foundation  may  consist  of  anything,  so  long  as  it  is  homogeneous 
and  stiff  enough  to  support  the  bedplate  at  all  points  when  the  pump  is  hi 
operation.  Without  proper  supports,  bedplates  cannot  be  expected  to 
maintain  the  proper  alignment  of  the  pump  shaft. 

Erection.  —  A  combined  bedplate  for  pump  and  motor  should  be  leveled 
up  by  wedges,  the  pump  and  motor  placed  upon  the  facing  strips  and  lined 
up  so  that  the  faces  of  the  pump  coupling  are  parallel,  and  the  pump  and 
motor  run  freely  with  and  without  coupling  bolts  in  position.  The  bed- 
plate should  then  be  grouted  into  place  so  that  it  is  absolutely  rigid. 
After  the  foundation  bolts  have  been  permanently  set  the  suction  and  dis- 
charge piping  may  be  connected. 

Suction  Piping.  —  Care  should  be  taken  to  see  that  there  are  no  air 
pockets  in  the  suction  piping  and  that  all  joints  are  absolutely  air-tight. 
The  pump  should  be  placed  as  near  as  possible  to  the  level  of  the  water  in 
the  suction  well,  but  in  cases  where  the  supply  is  distant  there  should  be  a 
continuous  rise  in  the  suction  pipe  toward  the  pump. 

The  suction  lift  or  vertical  distance  from  top  of  impeller  to  suction  water 
level  should  never  exceed  25  feet  at  sea  level,  20  feet  at  \  mile  above  sea 
level,  and  16  feet  at  1  mile  above  sea  level. 

Discharge  Piping.  —  The  discharge  piping  should  be  as  direct  as  possible 
and  with  a  minimum  number  of  bends.  A  tapered  section  should  be  placed 
at  the  pump  discharge  to  increase  the  diameter  of  the  delivery  to  the  same 
size  as  the  suction.  If  a  centrifugal  pump  has  too  much  resistance,  the 
whole  power  of  the  pump  may  be  expended  in  overcoming  the  frictional 
head,  and  give  no  discharge,  hence  the  necessity  of  giving  the  water  favor- 
able passage  during  its  travel  through  the  pipes  and  the  pump  and  avoiding 
all  unnecessary  resistance.  The  total  head,  theoretically,  which  a  centrif- 

F* 

ugal  pump  will  produce  is  ^- ,  where  V  is  the  peripheral  velocity  of  the 

impeller,  and  the  discharge  valve  is   entirely  closed.      The  peripheral 

27 


28  CENTRIFUGAL  PUMPING  MACHINERY 

velocity  required  to  support  a  column  of  water  of  given  height  is  equal  to 
the  velocity  acquired  by  a  body  when  it  has  fallen  through  a  height  equal 
to  that  of  the  column.  This  velocity  is  usually  expressed 

V  =  V2^H  or  V64.4  X  H, 

where  H  is  the  total  vertical  distance  in  feet  from  the  suction  level  to  that 
of  the  discharge.  All  friction"al  resistances  must  be  included  in  this.  To 
this  must  also  be  given  an  increment  in  order  to  effect  a  discharge  or  de- 
livery of  water.  In  carefully  designed  and  highly  efficient  pumps,  the 
periphery  velocity  is  found  to  be  less  than  that  given  above.  The  particu- 
lar equations  later  described  in  this  book,  were  developed  on  the  basis  of  a 
complicated  theory  obtained  from  the  characteristics  of  each  type  and  size 
of  pump  under  close  tests.  When  the  water  has  been  set  in  motion  the 
pump  will  continue  to  deliver  even  when  the  velocity  only  equals  A/2  gH. 

Lubrication  of  Bearings.  —  After  erection  and  before  starting,  all  pump 
bearings  should  be  carefully  washed  out  with  clean  petroleum.  It  is  most 
essential  that  the  oil  used  for  lubrication  should  be  of  the  best  quality  for 
high  speeds,  and  absolutely  free  from  all  trace  of  grit  or  sediment. 

Priming.  —  It  is  essential  to  have  a  gate  valve  at  the  point  of  discharge 
from  the  pump,  which  should  be  closed  before  the  pump  is  started.  The 
pump  should  then  be  primed  in  one  of  the  ways  described  in  Chapter  IV. 
When  the  pump  and  suction  pipe  are  filled  and  all  air  expelled,  the  air  cocks 
should  be  shut  and  the  various  pump  and  motor  bearings  carefully  examined 
before  the  motor  switch  is  closed.  Pumps  should  never  be  run  without 
water. 

Starting.  —  As  a  rule,  the  pump  will  come  up  to  normal  speed  in  about 
twenty  seconds,  after  which  the  gate  valve  on  the  discharge  should  be  very 
slowly  opened,  allowing  the  full  load  to  come  on  the  motor  as  gradually  as 
possible. 

Stuffing  Boxes.  —  The  packing  in  the  pump  stuffing  boxes  should  be 
carefully  adjusted  and  on  no  account  should  the  glands  be  screwed  too 
tight.  To  prevent  air  leakage  through  the  stuffing  box  on  the  suction  side 
of  the  pump,  a  gland  cage  may  be  provided  within  the  stuffing  box.  On 
each  side  of  the  cage  there  should  be  placed  about  three  rings  of  good 
hydraulic  graphite  packing.  A  J-inch  pipe  from  the  discharge  of  the  pump 
should  be  run  to  the  stuffing  box  and  an  opening  connected  through  to  the 
cage.  This  makes  a  water  seal  in  the  stuffing  box  and  prevents  all  air 
leakage.  The  gland  should  be  run  as  loose  as  possible,  otherwise  the  pack- 
ing is  liable  to  cut  the  shaft.  A  small  amount  of  leakage  from  the  stuffing 
box  does  no  harm,  in  fact  is  an  advantage,  as  it  prevents  the  packing  from 
heating  and  at  the  same  time  keeps  it  lubricated. 

Suction  Foot  Valve  and  Strainer.  —  In  all  centrifugal  pumps  where  there 
is  a  suction  lift,  it  is  absolutely  essential  that  a  foot  valve  of  proper  design 


OPERATING  29 

should  be  fitted  to  the  suction  pipe,  unless  special  means  are  provided  for 
emptying  the  pump  and  suction  pipe.  Careful  attention  should  be  given 
to  this  valve,  which  is  described  in  Chapter  IV.  The  foot  valve  should  in 
all  cases  be  fitted  with  a  strainer  having  an  area  of  not  less  than  twice  the 
cross-sectional  area  of  the  pipe. 

Before  starting  the  first  time  it  is  advisable  to  clean  out  thoroughly  the 
bearings,  including  the  thrust  bearing,  by  pouring  in  kerosene  and  allowing 
it  to  run  out  at  the  bottom,  as  dirt  is  liable  to  get  into  the  bearings  during 
shipment.  The  bearing  should  then  be  filled  as  full  as  possible  with  a  first- 
class  lubricating  oil  similar  to  dynamo  oil.  After  priming,  the  pump  can 
be  started  and  brought  up  to  speed  with  the  discharge  valve  closed.  The 
discharge  valve  can  then  be  opened  until  the  desired  quantity  of  water  is 
obtained.  If  the  total  head  operated  against  is  greater  than  that  for  which 
the  pump  was  designed,  the  quantity  of  water  discharged  will  be  less  than 
full  capacity,  and  there  may  even  be  no  discharge  at  all  if  the  head  is  suffi- 
ciently great,  but  the  power  will  be  less  than  that  required  by  the  designed 
head.  If  the  total  head  is  less  than  the  head  for  which  the  pump  is  de- 
signed, the  amount  of  water  discharged  will  be  greater  than  the  normal 
capacity  of  the  pump,  and  the  power  consumed  will  also  be  greater  than 
the  normal  rating. 

The  pumps  can  be  run  on  a  reduced  head  by  throttling  the  discharge 
until  the  desired  quantity  of  water  is  obtained.  This  creates  a  friction 
head  and  reduces  the  efficiency  of  pumping,  but  will  prevent  overloading 
the  motor. 

In  operation,  the  only  attention  which  the  pumps  require  is  an  occasional 
inspection  of  the  oil  in  the  bearings  and  of  the  packing  in  the  stuffing  boxes. 


PART   IT. 

CHAPTER  VIII. 

GENERAL  REMARKS. 

DISCUSSION  OF  THE  CENTRIFUGAL  PUMP. 

No  theory  of  the  centrifugal  pump  has  been  written.  In  making  this 
statement  the  word  "theory"  has  been  used  in  its  correct  sense,  which 
means  the  formulating  of  the  results  of  observation  into  laws  and  not  mere 
hypothesis.  A  clearer  idea  of  the  significance  of  the  word  "  theory  "  would 
cause  less  dispute  and  show  the  fallacy  of  the  common  expression,  "It  may 
be  right  in  theory  but  wrong  in  practice."  This  is  a  contradiction,  since 
the  theory  cannot  be  correct  if  it  does  not  fully  agree  with  practice.  Hy- 
pothesis is  really  meant,  and  the  statement  should  really  be,  "  The  hypothesis 
or  assumption  does  not  agree  with  facts."  Men  have  been  criticized  for 
being  theoretical,  and  fault  has  been  found  with  their  mathematics  when 
the  real  trouble  lay  in  a  misconception  of  the  theory  of  the  problem  and  a 
consequent  incorrect  application  of  mathematics. 

The  so-called  theories  of  centrifugal  pumps  are  analyses  based  on  certain 
assumptions.  Three  so-called  theories  will  be  presented  here  which  will 
give  approximately  the  same  results.  The  basis  of  the  principal  one  of 
these  is  that  the  motion  of  the  particles  of  water  throughout  the  space 
between  the  blades  strictly  follows  those  particles  contiguous  to  the  blades. 
Another  assumes  that  the  water  enters  the  inner  periphery  of  the  wheel 
in  a  radial  direction  across  the  entire  width  of  the  wheel.  Both  of  these 
are  wrong  and  are  made  only  because  the  problem  would  otherwise  present 
too  many  difficulties,  and  in  fact  be  impossible  of  solution.  There  are 
several  analyses  of  the  theory  of  centrifugal  pumps,  intricate  but  very 
useful  to  the  student,  and  authors  have  discussed  the  problem  in  a  masterly 
manner.  Such  writers  as  the  Russian  engineer,  T.  H.  Brix,  Dr.  Egon  R. 
v,  Greenebaum,  and  Engr.  Fritz  Neuman  have  given  complete  analyses. 
Unfortunately,  there  is  no  such  analysis  in  English.  The  formula?  given 
in  this  book  will  be  such  as  will  be  readily  available  for  the  designer. 

Wheels  usually  have  their  blades  made  in  the  form  of  an  arc  or  arcs  of 
circles.  An  exception  is  sometimes  found  in  commercial  standard  wheels 
where  the  outer  extremity  has  the  form  of  a  cycloid,  drawn  by  eye.  Some 
wheels  have  blades  of  the  form  of  an  Archimedian  spiral. 

31 


CHAPTER  IX. 

FIRST  THEORY  OR  ANALYSIS. 
THEORY  OF  IMPELLERS. 

THE  power  required  to  rotate  the  wheel  is  based  upon  the  change  in  the 
velocity  of  the  water  as  it  passes  from  the  inner  to  the  outer  periphery  of 
the  wheel  and  is  expressed  by  the  equation 

W  n7T7         i      WH 
—  (VU-vu)  =-fcT-> 
g  & 

in  which 

W—  the  weight  in  pounds  of  the  water  pumped  per  second; 
H  =  the  vertical  distance  in  feet  between  the  levels  of  the  water 

in  the  suction  and  discharge  reservoirs; 
v  and  V  =  the  tangential  velocities  of  the  wheel  in  feet  per  second  at  its 

inner  and  outer  periphery  respectively; 
u  and  U  =  the  tangential  velocities  in  feet  per  second  of  the  water  at 

the  inner  and  outer  periphery  of  the  wheel  respectively; 
g  =  the  acceleration  due  to  gravity ; 
E  =  the  hydraulic  efficiency  of  the  wheel. 

The  left-hand  member  of  the  above  equation  represents  the  work  re- 
quired to  overcome  the  hydraulic  resistance  of  the  wheel,  but  does  not  include 
any  mechanical  resistances  such  as  friction  of  the  bearings,  etc.  The  right- 
hand  member  represents  the  work  done. 

This  equation  contains  the  fundamental  principle  used  in  calculating  or- 
dinary centrifugal  pumps  as  given  by  most  of  the  writers.  As  stated,  it  is 
assumed  in  ordinary  calculations  for  centrifugal  pumps  that  the  water  enters 
the  eye  of  the  wheel  radially,  and  consequently  the  term  vu  becomes  zero, 

and  we  may  write  V  =  Jriy  •    From  the  triangle  of  velocities  in  Fig.  28  it 
will  be  easy  to  deduce  the  following  formulae : 

/gH(sm(A-~B)\     JgH (       tangZA         /gH  /        tang~B\ 
"  V  E  \smAcosBj      V  E  \       tangA/      V  E  \     r  tang  D/ 

IgH/  1  +  cotT\ 
V  E  \    ^  cot  CJ 

tangg  =  tangg  =  cot  A  =       cot  T  EV2 

tang  A      tang  D      cot  B  ~      cot  C  ~      "  SH  ' 
32 


FIRST   THEORY  OR  ANALYSIS 


33 


and 


•j  =  cot  B  -  cot  A  =  tang  T  +  tang  C, 


cot  B  = 


The  above  embrace  all  the  formulae  given  hi  various  forms.  The  greater 
the  number  of  blades  the  more  nearly  will  the  assumption  agree  with  the 
facts,  but  too  many  blades  will  cause  undue  friction.  Brix  and  other  writers 
give  complicated  formulae  for  the  number  of  blades,  but  this  is  usually 


< Tang.Vel.  U    ofWi 


Cot.D 


E  = 


V  (Y- 


Fig.  28.     Triangle  of  Velocities. 


assumed.  The  diameter,  speed,  and  width  of  the  wheel  are  dependent  on 
the  motor  and  on  the  allowable  velocity  of  the  wheel  and  the  water  through 
it,  features  which  each  designer  will  have  to  settle. 

The  angle  of  the  blade  as  well  as  its  radius  is  determined  by  the  following 
geometrical  and  analytical  considerations.  To  construct  the  curve  of  the 
blade,  it  is  necessary  to  know  the  inner  and  outer  radii  of  the  wheel,  and  the 
angles  <£  and  D  which  the  blade  makes  with  the  inner  and  outer  periphery 
of  the  wheel.  The  angle  D  is  determined  by  calculation  from  formulae 
given,  and  usually  the  radial  velocity  at  the  inner  periphery  of  the  wheel  is 
assumed  to  be  the  same  as  that  at  the  outer;  the  angle  </>,  unit  of  blade, 
may  be  graphically  determined  as  shown  in  Fig.  29. 


34 


CENTRIFUGAL  PUMPING  MACHINERY 


Construction  of  blade 
Given  R,r,  D,and$ 


Referring  to  Fig.  29  we  proceed  as  follows:  Let  r  and  R  be  the  radii  of 
the  wheel.  From  a  draw  a  line,  making  the  angle  D  as  shown,  and  a  line 
from  the  center  0,  making  the  angle  D  +  <£.  From  a  draw  the  line  abc] 

then  c  will  be  the  inner  extremity 
of  the  blade.  Draw  the  line  de 
perpendicular  to  and  bisecting 
the  line  ac  ;  then  the  point  d  will 
be  the  center  of  the  curve  of  the 
blade  as  shown.  This  assumes 
that  the  angles  <£  and  D  are 
known. 

To  find  the  angle  4>,  the  veloc- 
ity v  at  the  inner  circumference 
of  the  wheel  varies  inversely  with 
ratio  of  the  radii  and  directly 
with  the  velocity  of  the  wheel  at 
the  outer  circumference,  that  is, 


Fig.  29.     Construction  of  Curve  of  Blade. 


Since  it  has  been  assumed  that 
the  water  enters  the  eye  of  the 
wheel  radially,  it  has  no  velocity 
of  rotation,  and  the  triangle  of 
velocities  is  as  shown  in  Fig.  30. 
If  Q  be  the  volume  of  water  in  cubic  feet  pumped  per  second,  w  the  width 

of  the  wheel  at  the  outer  circumference,  then,  disregarding  the  thickness  of 

the  blades,  the  radial  velocity  of 


the  water  at  the  outer  circumfer- 
ence of  the  wheel  will  be 

Q 


Determinate  n  of  Angle  d> 


2irrw' 

all  dimensions  being  in  feet. 

The  relative  velocity  of  the 
water  at  entrance  and  exit  of  the 
wheel  should  have  the  same 
direction  as  that  of  the  blade 
at  the  corresponding  points,  and 
this  is  what  governs  the  angles 
D  and  <£.  If  this  is  not  followed 
injurious  shocks  will  result. 

This  discussion  does  not  consider  the  influence  of  a  vortex  chamber  or 
the  form  of  the  discharge  passage  around  the  wheel. 


ab 


Fig.  30.     Triangle  of  Velocity  with  Water 
Entering  Wheel  Radially. 


FIRST   THEORY   OR   ANALYSIS 


35 


The  calculation  of  an  8-inch  three-stage  pump,  with  details  of  impeller 
and  diffusion  vanes,  according  to  the  above  analysis,  is  shown  hi  Fig.  31 
and  described  below. 


Impellers  and  Diffusion  Rings 


Fig.  31.     Impeller  and  Diffusion  Vanes  of  8-inch  3-stage. 


CALCULATION  OF  8-INCH  3-STAGE 

Conditions:       1400  gallons  per  minute. 

1170  revolutions  per  minute. 
210  pounds  pressure  or  70  pounds, 
each  stage. 

Outer  diameter  of  impeller 18 Jf" 

Inner  diameter  of  impeller 8|" 

Hub  diameter  of  impeller 4f " 

Width  of  opening  at  outer  circumference 

of  impeller 1" 

Width  of  opening  at  inner  circumference 

of  impeller 2f" 

Number  of  blades 12 

Thickness  of  blades  at  outer  circumference  f  " 

Outer  diameter  of  diffusion  ring 27" 

Inner  diameter  of  diffusion  ring 18%" 

Number  of  guides  in  diffusion  ring 10 


36 


CENTRIFUGAL  PUMPING  MACHINERY 

Formula  Used. 


JS  .  tan*,  +  tanft. 


DttfoBion  Bing 


vf 


J 
sfficiency. 


tan  02  / 
1  g  H  cot  <t>z 


VzU^  tan  </>2  vtfiz 

H  =  net  or  useful,  here  70  pounds  head. 
All  dimensions  in  feet;  velocity  in  feet  per  second. 


log  rx  =  log  r2  +  cot  fa  X  ~  X  log  e. 


180 

For  equal  steps  of  4>°  there  is  a  (  Trcotfo  log  e  I  s  ^  _  ^ 
constant  difference,  for  log  rx  =    (     180  j 

These  dimensions  in  inches. 

18.46875  X  1170X7T 

12X60 

d2  -  18.46875      log  1 .  2664374 

TT       log  0.  4971499 

1170       log  3. 0681859 

4. 8317732 

60  X  12  =  720      log  2.8573325 
vz  =  94.  284  ft.     log  1.  9744407 


18.  46875  log  1.2664374 

TT        0. 4971499 

1.7635873 


Outer  circumference  of  impeller  =  58.  02  in. 
Blades  12  X  I  =  4-  50  m- 

Net  outer  circumference  53.  52  in. 

As  opening  is  1  inch  wide  this  is  also  the  area  of  the  opening  in  square  inches. 
1400  X  231 
60 


5390  log  3.  7315888 

53.52      log  1.7285161 

12.00      log  1.0791812 

2.8076973 

8. 3925  log  0.  9238915 


r2  = 
cos  52  =  8 

log  554 
591 


+  9H  =  554 
591 

2.7435098 
2.7715875 


or 


log  cos  of  20°-22'-53.  55"  =  9.  9719223 
cos  52  =  sin  02  tan  02  =  cot  52. 

52  =  20°-20'-53.  55"        log  cot     0.  4307806 
tan  of  02  =  tan69°-39'-6.45" 

v,  =  94.284    log  1.9744407 
uz        8.  3925  log  0.  9238915 

^      11.234    log  1.0505492 


tan  nat.  02 
tan  nat.  <£2 


2.692 
8.  542 


FIRST    THEORY   OR   ANALYSIS 


37 


=  830-19'-22.  28"     tan  02  =    2.692      log  0.  4307806 

tan  02  =    8. 542      log  0.  9315596 

tan  02  -J-  tan  02  =    0. 31566  log  9. 4992210 

0  =  32.16        log  1.5073160 

H  =  70  Ibs.  =  161 .  63  ft.  log  2.  2084414 

3. 7157574 

t>22  log  3.9488814 

9. 7668760-10 
log  0.1 191007-10 


tan02 
Efficiency  17 


1.31566 
0. 76908 


log  9.  8859767 


This  is  the  hydraulic  efficiency  as  computed  from  the  formula. 


Diffusion  Ring. 

Calculation  of  one  radius  vector  of  the  curve  of  guide. 

<t>°  X 


log  TX  =  log  r2  +  cot  02  X 


180 


Xloge. 


re0°  =  36°;  then 


log  rx  =  log  9.25  +  cot  02  X  £  X  log  e 
o 

cot  02  log  9. 0684404-10 

TT  log  0.4971499 
log  e  log  9.  6377843-10 
9. 2033696-10 
5  log  0. 6989700 
0. 0319447  log  8.  5043996-10 
log  9. 25  0.9661417 

^  cot  02  log  e     0. 0319447 
0. 9980864 

rx  =  9.956" 
r2  =  9.25" 
diff.  =0.706" 

2 


Gain  in  pressure  in  free  vortex  =  ~-  j  1  —  ( 


t*  — 


uz  u2  =  8. 3925  log  0.  9238915 

tan  52  tan  52      9. 5692194^-10 

22.629  9.3546721-10 

v2  =  94.284 
-f-  tanS2  =  22.629 


u, 


tan  62 


71.655 

log  71. 655  or  ^  =  1. 8552464 
log  o>22        =  3.  7104928 
V   1547 
j   2916 


1547  log 
2916  log 


3. 1894903 
3. 4647875 


9. 7247028-10 
2. 9  =  64. 32  log  1.8083460 
0. 0082481   log  7.  9163568-10 


-6T 


3. 7104928 
7. 9163568-10 


1.6268496 
42. 349  ft.  gain  in  head,  or  18. 34  pounds. 


38 


CENTRIFUGAL  PUMPING  MACHINERY 


To  set  this  before  the  reader  in  simple  form,  the  following  diagrams,  with 
calculation,  have  been  developed,  which  can  be  applied  to  impellers  and 
diffusion  vanes  in  turbine  pumps  (Fig.  32). 

STUDY  OF  CENTRIFUGAL  PUMP. 


'     k 

q 

\    1 


Graphical  Solution. 
Ha  =  useful  lift  in  feet. 
77  <  1  =  efficiency. 
All  dimensions  in  feet. 
All  velocities  in  feet  per  second. 


Radial  velocity  =  cubic  feet  per  second  divided  by  the  net 
circumferential  area  in  square  feet;  that  is,  allowance  must 
be  made  for  thickness  of  blades. 

Fig.  32.     Diagram  for  Calculating  Impellers  and 
Diffusion  Vanes. 


=  y22c22(l  -  sin2  02) 
=  t'22c22  cos2  02 

sin2 


4 


cos 


sin2  (at  - 


. 
' 


hence, 


tan  02  _  1       tan  02 
tan  a.-!  ~        tan  <f>2 


ua 


Z-L  = 


COS  02  COS  02 

V2  =  c2  sin  02  +  z2  sin  02 
_       sin  02  sin  02 

2  cos  02         2  cos  02 
=  Uz  tan  02  -}-  u<i  tan  02, 

—  =  tan  02  +  tan  02. 

Curve  of  Guides  in  Whirlpool  Chamber. 

rx  =  r2e*  cot  ^2  (0  in  circular  measure) 
0°X-n- 
180 


(2) 


loge 


FIRST    THEORY   OR   ANALYSIS 


39 


Or  we  may  calculate  one  value,  such  as  n,  and  construct  all  others.  The  vector  bisecting 
the  angle  between  any  two  vectors  is  a  mean  proportional  between  these  two  vectors. 
Thus,  r32  =  r2r3,  a  property  of  the  logarithmic  spiral.  Describe  semicircle  bac,  erect 
perpendicular  Oa,  draw  arc  ad  from  0  as  center,  and  d  gives  point  in  curve.  Continued 
bisections  will  give  new  points. 

To  draw  the  blades:  Lay  off  angle  81  +  S2,  as  shown,  and  line  through  /  gives  h,  the 
inner  end  of  blade.  Draw  line  5  at  an  angle  S2,  as  shown,  and  the  point  p  at  the  inter- 
section of  line  perpendicular  to  and  bisecting  line  hf  is  the  center  of  the  curve  of  blade. 

THEORY  OF  DIFFUSION  GUIDES  AND  VANES. 

The  calculation  of  the  curve  of  the  diffusor  guides  or  vanes  in  the  cham- 
ber is  shown  in  Fig.  33.  Let  r  be  the  distance  from  the  center  of  the 
wheel  to  any  point  of  the  curve,  and  B  the 
angle  which  the  tangent  to  the  curve  at  that 
point  makes  with  radius,  which  is  constant  at 
all  points,  this  being  a  property  of  the  logarith- 
mic spiral.  This  angle  B  is  also  that  which 
the  absolute  velocity  of  the  water  as  it  leaves 
the  wheel  makes  with  the  tangent  to  the  wheel 
at  the  outer  circumference.  We  then  have  as 
the  equation  of  the  path  of  the  water  in  the  Fi8-  33-  Velocity  Diagram  for 
chamber,  Diffusion  Vanes. 

log  r  =  log  a  +  cot  B  X  </>  X  log  e, 

where  a  is  the  radius  of  the  outside  of  the  wheel,  <j>  is  the  angle  in  circular 
measure  which  the  radius  through  the  point  in  the  curve  makes  with  the 
radius  of  the  point  at  the  outer  circumference  of  the  wheel,  and  e  is  the 
Napierian  base. 

If  <t>°  is  the  angle  in  degrees,  then  <t>  =  <t>°  X  3.1416  -r-  180. 
Example.  — 

a  =  9.25  inches; 
B  =  80°; 

</>  =  n  X  10°  X  3.1416  -4-  180, 
where  n  is  a  multiple  of  an  angle  of  10°.     We  then  have 

log  r  =  log  9.25  -h  (cot  80  X  n  X  10  X  3.1416  X  log  e  -4-  180). 


cot  80  log  9.2463188  -  10 

cot  10  log  1.0000000 

3.1416  log  0.4971499 

log  e  log  9.6377843  -  10 

0.3812530 
180  log  2.2552725 

0.0133654  log  8. 1259805  -  10 
9.25  log  0.9661417 

0=    (log  r  -  log  a)  X  180 
3.1416  X  cot  B  X  loge 


log  r  =  0.9661417  +  0.0133654  X  n. 
If  0°  =  30°,  then  n  =  3  and 
logr  =  0.9661417  +  0.0400962 
=  1.0062379  and  r  =  10.14467. 

Suppose  r  is  given  and  we  wish  to 
find  the  corresponding  angle  <f>°,  then 
we  have 

r  =  10.14467  log  1.0062379 
a  =  9.25         log  0.9661417 


0.0400961 


40 


CENTRIFUGAL  PUMPING  MACHINERY 


0.0400961  log  8.6031020 
180     log  2.2552725 


10 


0.8583745 
9.3812530  -  10 


3.1416 
cot  80° 
loge 


log  0.4971499 

log  9.2463188  -  10 

log  9.6377843  -  10 

9.3812530  -  10 


<t>°  =  30.00001  log  1.4771215  or  30°,  which  checks 
0°  =  30°  =  10  X  n  =  10  X  3. 

Fig.  34  will  illustrate  graphically  this  method  applied  to  example. 


Fig.  34.     Graphical  Method  for  Diffusion  Vanes, 
log  r  =  log  a  +  cot  B  X  0  X  log  e,  given  0° 

*  =  ^wL 

0  =  (log  r  -  log  a)  X  180°    0°  unknown 
TT  X  cot  B  X  log  e          given  r 

As  frequently  stated,  the  object  is  to  utilize  the  velocity  of  the  water  as 
it  issues  from  the  wheel  by  changing  the  energy  of  motion  into  energy  of 
pressure.  To  obtain  the  full  benefit  of  a  vortex  chamber,  the  direction  of 
flow  of  the  water  must  always  be  the  same;  in  other  words,  the  angle  which 
the  water  path  at  any  point  makes  with  the  corresponding  radius  must  be 
a  constant.  Furthermore,  the  line  of  direction  must  be  that  of  the  absolute 
velocity  of  the  water  as  it  leaves  the  wheel.  The  equation  to  this  curve, 
which  is  a  logarithmic,  spiral  is  r  =  aem<t>  when  0  =  0.  e  =  Napierian  base, 
r  =  a.  m  is  the  point  determined  by  calculation  from  the  problem.  In 
the  ordinary  type  of  turbine  pump  the  water  cannot  take  this  path,  except 


FIRST    THEORY  OR  ANALYSIS 


41 


the  last  wheel,  but  must  flow  around  the  circumference  of  the  wheel. 
An  easy  turn  is  made  from  the  correct  curve.  The  area  of  the  passageways 
at  the  circumference  is  equal  to  the  area  of  discharge  pipe,  or  slightly  larger, 
that  there  may  be  no  sudden  velocity  changes.  In  some  recent  pumps  the 
flow  of  water  throughout  the  pump  from  suction  to  discharge  is  in  a  con- 
tinuous spiral  path  in  order  to  obviate  all  shocks  and  sudden  changes. 
The  curves  depend,  however,  upon  the  fact  that  the  radial  flow  in  the  dif- 
fusion ring  varies  inversely  as  the  distance  from  the  center,  and  the  curve 
of  the  path  is  the  resultant  of  the  absolute  and  radial  flow.  A  method  of 
constructing  diffusion  rings  by  aid  of  circles,  not  governed  by  any  laws,  is 
shown  in  Fig.  35.  The  logarithmic  spiral  curve  will  give  the  best  efficiency. 
A  simple  way  to  construct  curves  in  the  diffusion  ring  is  shown  in  Fig.  36. 
Let  rx  be  any  radius  of  the  curve  calculated  by  rx  =  Ree  «*  *,  where  0  in 
0°X  3.1416 


circular  measure  is  = 


180C 


Bisect  angle  0  by  line  o°  and  prolong.  The  bisector  cb  is  the  mean 
proportional  of  the  radii  R  and  rx.  Erect  perpendicular  od.  Describe 
arc  cdb  on  cb  as  a  diameter;  then  od,  cut  off  by  the  arc,  is  equal  to  the 
radius  of. 

a 

By  bisecting  -  we  obtain  another  point  in  a  similar  manner,  and  thus  as 

ft 

many  points  as  are  desired  may  be  found. 


NXArea=Discharge  Pipe 
N=No.  of  Vanes 


Fig.  35.  Fig.  36. 

Method  of  Constructing  Diffusion  Vanes  by  Circles. 


APPLICATION    OF    THEORY. 

In  order  to  cover  the  details  involved  in  the  mathematical  consideration 
of  the  problem  we  will  calculate  a  concrete  case,  and  will  consider  a  pump 
having  12  stages  with  approximately  50  feet  to  the  stage.  It  is  now  possible 


42  CENTRIFUGAL  PUMPING   MACHINERY 

to  work  single  stages  as  high  as  150  feet  and  even  more  under  proper  con- 
ditions, but  we  make  the  above  assumption  in  order  to  illustrate 

Pump,  4-inch,  12-stage. 
Capacity,  150  gallons  per  minute. 
Speed,  1460  revolutions. 
Total  or  useful  head,  610  feet. 
Diameter  of  inlet  and  outlet,  4  inches. 
Outside  diameter  of  diffusion  ring,  13  inches. 
Inside  diameter  of  impeller,  3J  inches. 
Diameter  of  hub  of  impeller,  2  inches. 
Outside  opening  of  diffusion  ring,  J-inch  wide  (assumed). 
Outside  diameter  of  impeller,  8jf^  inches,  to  suit  speed. 
Width  of  opening  of  impeller  at  outside,  f  inch. 
Inside  diameter  of  diffusion  ring,  9  inches. 

Angle  which  tangent  to  blade  makes  with  radius  of  outside,  or  02  =  72°. 
Number  of  blades,  12. 

Calculations.  —  Tangential  velocity  of  outer  circumference  of  impeller 
_  1460  X  8§  J>  inches  X  IT  _  146  X  287  X  TT 
60  X12  2304 

146  log        2.1643529 

287  log        2.4578819 

TT  log        0.4971499 


5.1193847 
2304  log        3.3624825 


v2  =  57.13500         1.7569022  feet  per  second. 
Radial  velocity  of  water  at  outside, 

150  X  231 


1 


60 


i  X  f  X  TT  -  12  X  t  X  f 
The  thickness  t  of  the  blades  at  the  outer  circumference  measures  f  ,  and 
12X|  =  4.5. 
The  circumference  of  8§|  =  28.676; 

of  28.676  -  4.5  =  23.676,  leaving  a  net  length,  then 
150  X  231 

60  385  . 


12      7L028        ' 
?*  =  5-^-  =  10.56  and  -  =  10.56  =  tang  02  +  tan  02 

Uz  O.4Z  U-2. 

tang  62  =  tang  72°  =  3.0777 

tang  0  =  10.56  -  3.0777  =  7.4823, 


FIRST    THEORY  OR  ANALYSIS  43 


we  have 


gH 

gH                     32.16 

X610 

u  tang  fa 

tang  fa  = 

n  — 

77      ?/2t;2tang4>2      12  X  u2  X  vz  X  tang<£2 

32.16  log      1.5073160 
610      log     2.7853298 

4.2926458 
12          log      1.0791812 
5.42      log     0.7339993 
57.135    log      1.7569022 
7.4823  log     0.8740351 

er  cent  effi( 

4.4441178 
0.70555          1.8485280=  70  p 

Path  of  the  water  or  curve  in  diffusion  ring, 

rx  =  radial  distance  to  any  point 

=  rx  =  re  cot  fa  X  <f>, 

or  log  rx  =  log  r  +  <£  X  cot  fa  X  loge, 

where  r  —  inside  radius  of  impeller  =  4J  inches. 

0  =  circular  measure  of  angle  between  r  and  rx  (see  Fig.  37)  ,  in  de- 


, 
grees  to 


e  =  Napierian  base. 
rx  is  very  simply  calculated  : 

(1)  rxl  =  log  r  +  cot  fa  X  0  X  loge. 

(2)  rx    =  log  r  +  cot  faX<t>'x  loge. 

(3)  rx"  =  log  r  +  cot  <£2  X  <£"  X  loge. 

The  difference  between  (2)  and  (1)  and  (3)  and  (2)  is 

(</>'-<£)  (Cot  fa  loge), 

and  by  making  rx  progress  by  uniform  steps  this  is  a  constant. 
Make  the  difference  in  angles  5°;  then 

_5XxXir_Xir 
HLSOT    ~36' 

where  x  =  the  multiple  of  5  in  the  angle  0.     Then 

XTT 
rx  =  logr  +  cot<£2X       Xloge 


XTT 
log  4.5  +  cot  fa  X       X  log  e 


44 


CENTRIFUGAL  PUMPING  MACHINERY 


Note  where  the  diffusion  ring  is  used  02  is  not  so  important  as  without. 
tang  02  =  7.4823  log  0.8740351 


cot  02  = 


log  1.1259649 


.        ^ 
tang  02 

log  cot  02  =    .1259649 
TT  0.4971499 

log  of  log  e   1.6377843 


log  36 


1.2608991 
1.5563025 


=  0.0050652  log  3.7045966 
log  4.5    0.6532125 


0  =  0 
5° 
10° 
15° 
20° 
25° 
30° 
35° 
40° 
45° 
50° 
55° 


Constant  0.0050652  rx 

rx  =  r  log  0.6532125  4.5 

rx  =  log  0.6582777  4.552 

0.6633429  4.606 

0.6684081  4.660 

0.6734733  4.715 

0.6785385  4.763 

0.6836037  4.826 

0.6886689  4.883 

0.6937341  4.940 

0.6987993  4.999 

0.7038645  5.057 

0.7089297  5.116 
We  have  simply  to  add  constant  above 
to  log  of  r. 

To  obtain  rx  =  for  0°  =  50°  we  have 
added  the  constant  ten  times: 

10  X  0.0050652 

=  0.0506520 

log  r  =  0.6532125 

log  rx  for  0°  =  50    0.7038645 

This  checks  the  work. 

The  next  step  is  to  find  the  arc  of  a 
circle  which  shall  pass  through  the 
greatest  number  of  points  as  given  by 
rx  (see  Fig.  38).  The  center  of  this  curve 
should  be  on  a  line  perpendicular  to  the 
tan  at  r  =  4.5,  or  very  nearly  so. 

Regarding  the  diffusion  ring  we  may 
make  the  following  observations:  The 
width,  or  difference  between  the  inner 
and  outer  radii,  varies  with  different 
builders,  as  no  one  can  tell  the  exact 


jrjgs  37^  35  an(j  39 
Method  of  Designing  Diffusion  Vane. 


FIRST   THEORY   OR  ANALYSIS 


45 


gain  in  passing  through  the  ring.  It  will  not  do  to  have  too  large  an  angle 
02  or  too  many  guides,  as  the  space  becomes  so  small  that  the  velocity  is  very 
great,  and  it  should  not  be  more  than  the  absolute  velocity  of  the  water 
leaving  the  wheel.  Where  the  water  discharges  around  the  ring  the  curves 
at  the  end  should  be  radial  as  shown  in  Fig.  39.  The  angle  ^  is  easily 
calculated,  thus  I X  w  when  w  =  width  is  found;  then 

d*  * 

TT —  =  ratio  of  total  opening, 
Dwir 

with  reference  to  discharge  pipe.    Here 

D  =  13;    w  =  \  inch;    d  =  4  inches; 

16 
4 


then 


13 


16       2        8^ 
4       13      13' 


X  360°  = 
lo 


=  221.50 


We  find  6  guides  give  about 
the  right  passage  for  the  veloc- 
ity and  space  at  the  outer  cir- 
cumference ;  we  have,  therefore, 

=  36.9°.    We  will  call  it, 


Figs.  40  and  41. 
Method  of  Designing  Impeller  Blade. 


Fig.  42. 
Section  of  Impeller. 


for  practical  purposes,  35°,  and  then  round  the  corners.  The  water  will 
j  then  enter  the  passage  around  the  wheel  at  about  the  same  velocity  as  the 
[water  in  the  passage.  When  this  is  not  the  case  there  should  be  a  free 
!  passage  on  an  Archimedean  spiral. 


46  CENTRIFUGAL  PUMPING  MACHINERY 


BLADES. 

A  few  remarks  only  are  necessary.  To  be  sure  of  the  inner  and  outer 
angles  the  center  0,  Fig.  40,  should  be  such  as  to  give  a  short  straight  line 
at  both  ends,  a  and  C.  Strictly  speaking,  the  water  should  enter  the  im- 
peller in  the  direction  of  the  blade  at  the  inner  end,  and  as  it  is  usual  to 
make  the  radial  velocity  the  same  at  the  inner  and  outer  circumference 
this  angle  may  be  determined  as  per  Fig.  41 ;  but  owing  to  the  doubt 
about  the  real  direction  of  entrance,  this  angle  is  usually  made  somewhat 
larger.  This  is  wholly  a  matter  of  judgment,  as  no  test  can  absolutely  de- 
termine it.  Much  has  been' written  on  it,  and,  strictly  speaking,  the  blade, 
having  the  section  A,  Fig.  42,  should  be  calculated  for  each  radius,  and 
formula  have  been  given  for  this.  But  as  the  calculation  is  complicated 
and  founded  wholly  upon  assumptions,  it  is  questionable  whether  it  pays  to 
consider  this.  In  fact,  most  authors  assume  the  tangential  velocity  of  the 
water  at  the  entrance  as  zero. 


CAPACITY    OF  CENTRIFUGAL  PUMPS. 

The  velocity  of  flow  through  the  impeller  governs  the  capacity.  When 
this  velocity  and  the  circumferential  area,  deducting  the  thickness  of  vanes, 
are  known,  the  total  capacity  can  easily  be  calculated.  The  speed  for  a 
given  discharge  can  be  calculated  and  depends  upon  the  head  against  which 
the  pump  is  to  work.  The  water  is  supposed  to  rotate  in  the  pump  as  a 
solid  mass,  and  delivery  commences  when  the  centrifugal  force  is  greater 
than  the  total  lift  including  all  losses  and  friction.  Let 

HI  =  absolute  velocity  inlet, 
Uz  =  absolute  velocity  outlet. 

u<?  —  Ui2 

Then  the  centrifugal  force  =  — ~ This  speed  is  the  maximum,  and  in 

*Q 

practice  it  will  be  less,  depending  on  the  angle  between  the  impeller  vane 
and  the  circumference.  The  efficiency  of  the  pump  is  also  dependent  upon 
this  angle.  It  should  be  noted  that  the  tip  angle  has  considerable  influ- 
.ence  upon  both  the  efficiency  and  the  uniformity  of  power  required. 

In  the  formula  known  as  Appold's,  v  =  550  +  500  y/Hf,  Hf  is  the  static 
head  in  feet,  =  nCf,  in  which  Cf  is  the  circumference  of  impeller  in  feet. 
500  is  an  arbitrary  figure  but  supposed  to  be  equal  to  v  2  gh  times  a  con-  • 
stant,  and  the  550  another  arbitrary  figure,  to  give  the  necessary  velocity 
in  feet  per  minute.     This  will  have  to  be  revised  in  order  to  meet  later 
developments  in  this  class  of  pumps.     The  entire  head  produced  by  ai 
pump  depends  upon  the  velocity,  disregarding  the  angles  of  the  vanes.     If' 
this  velocity  =  V,  and  total  head  =  H,  H v  =  velocity  head,  Hf  =  friction . 


FIRST    THEORY  OR   ANALYSIS  47 

head,  H,  =  static  head,  and  Vt  =  theoretical  velocity  of  H,  the  equation  then 
becomes 


On  closing  discharge  valve  H  will  equal  Ha; 


on  opening  the  discharge,    .V  =  -TT-  V2  g \HB  +  Hv  +  H/\ . 

v  t 

It  must  not  be  forgotten  that  the  radial  velocity  of  the  water  depends 
upon  the  dimension  of  the  wheel,  and  that  if  the  quantity  of  water  pumped 
is  excessive  and  the  radial  velocity  be  increased  the  frictional  losses  will 
also  be  increased  and  the  efficiency  diminished.  Some  manufacturers,  after 
establishing  their  sizes  and  details  of  design,  use  arbitrary  figures  for 
the  velocity  in  order  to  obtain  the  static  head;  i.e.,  for  #  =  static  head 
they  use  a  velocity  of  the  outer  circumference,  V  =  10  VH.  Where 
to  include  friction  losses  in  suction  and  delivery  they  use  V=  9 

V2 

When  discharge  valve  is  closed  the  pressure  produced  is  H  =  — .  In  a  cen- 
trifugal pump  the  power  is  directly  dependent  upon  the  capacity  at  a 
stated  speed.  The  total  head  is  also  dependent  upon  the  capacity,  there- 
fore the  head  created  can  be  used  as  a  useful  head  or  lost  by  closing  the 
valve  on  the  discharge.  A  partially  closed  valve  causes  a  lower  efficiency, 
which  bears  a  proportion  to  the  rated  efficiency  equal  to  the  ratio  of  the 
head  generated  to  the  total  head  generated.  Assuming  a  pump  working 
against  a  head  of  250  feet  and  giving  1000  gallons  capacity  at  75  per  cent 
efficiency,  what  would  the  efficiency  be  for  the  same  capacity  at  175  feet 
head?  The  probable  efficiency  will  be  75  X  H£  =  52.5  per  cent,  or  in 
other  words  the  horse  power  can  be  calculated  for  the  capacity  against 
175  feet  and  divided  by  the  brake  horse  power  at  250  feet. 

The  following  empirical  formulae  for  ready  and  approximate  figuring 
may  be  used: 

1830  X  VH  ,  .        „      .    .     , 

— ; — : —  — : —  -  =  diameter  of  impeller  in  inches, 

revolutions  per  minute 

This  will  apply  to  volute  pumps. 

For  turbine  pumps  with  diffusion  ring,  the  formulae  with  another  con- 
stant can  be  used : 

1850  VH  . . 

— : =  diameter  of  impeller  in  inches, 

revolutions  per  minute 

for  all  impellers  below  6  inches.  For  6  inches  and  larger  add  £  inch  to  the 
diameter  obtained,  and  for  every  3  inches  of  increase  in  diameter  add 
another  J  inch. 

For  dredging  work  the  formula}  become 

2000  VH 

— r— -. —          — : —  -  =  diameter  of  impeller  in  inches, 
revolutions  per  minute 


CHAPTER  X. 

SECOND   ANALYSIS   OR  THEORY. 
THEORY   OF  IMPELLERS. 

THIS  section  will  treat  the  subject  in  a  slightly  different  way  to  give 
the  designer  a  method  which  may  be  easier  and  can  be  more  readily 
applied. 

H  =  head  in  feet; 

Q  =  capacity  in  gallons  per  second  or  cubic  feet  per  second; 
n  =  revolutions  per  minute ; 
u2  =  circumferential  speed  in  feet  per  second; 
v  =  radial  speed  in  feet  per  second; 
p  =  power  in  horse  power  or  foot  pounds  per  second. 
a,  (3,  7  are  the  constants  in  the  general  equation,  used  to  show  the  rela- 
tion between  head,  speed,  and  capacity,  or  between  H,  u,  and  Q. 

r,  s,  t  are  the  constants  used  in  the  general  equation,  which  will  be  used 
to  show  the  relation  between  horse  power,  revolutions,  and  capacity,  or 
between  P,  n,  and  Q. 
g  =  acceleration. 

In  considering  the  following,  it  is  well  to  understand  that  all  curves  in  the 
figures  submitted  are  from  actual  performances,  and  details  have  been 
obtained  from  working  impellers  in  order  to  present  reliable  data  on  which 
to  base  conclusions.  Note  that  in  the  curves  submitted  the  abscissas 
show  the  capacity  in  gallons,  and  that  the  ordinates  give,  the  head  in  feet, 
from  which  we  get,  with  a  constant  speed,  what  is  termed  the  capacity- 
head  curve.  This  relation  between  capacity,  head,  and  speed  is  expressed 
in  the  following  equation,  which  is  that  of  a  hyperbolic  paraboloid  corre- 
sponding to  the  path  of  the  water  through  the  pump.  The  head  due  to 
the  centrifugal  force  will  vary  with  the  square  of  the  distance  from  the 
center,  and  the  curve  assumed  by  the  surface  of  the  water  will  be  that  of  a 
parabola, 

aQ2  -  PQn-yn2  =-2gh, 

Q,  n,  and  h  being  variables,  and  a,  /3,  7,  #,  constants. 

The  relations  existing  show  that  the  capacity  varies  directly  with  the 
speed,  the  head  being  constant,  that  the  head  will  vary  directly  as  the  square 
of  the  speed  at  constant  capacity,  and  that  the  head  will  vary  directly  as 
the  square  of  the  capacity  for  constant  speed.  These  relations  govern  all 
cases  and  must  be  clearly  understood  in  designing  and  operating  centrifugal 

48 


SECOND   ANALYSIS   OR   THEORY 


49 


s| 

5  uoc 


100     120     140      160     180 

Gallons  per  Minute 

Constant  Head. 


Fig.  43. 


Characteristics  Showing  Capacity  Relation  for 
Constant  Head. 


pumps,  and  in  selecting  pumps  suitable  for  a  particular  purpose.     Various 

curves  are  illustrated  for  analyzing  the  action  of  the  pump. 
Fig.   43   shows  the 

relations  for  constant 

head. 

Fig.   44  shows  the 

relations  for  constant 

speed. 

Fig.   45   shows  the 

relations  for  constant 

speed. 

Fig.  46  shows  curves 

for     variable     speed, 

with  the  equations  for 

capacity  -  head    curve 

and     capacity  -  power 

curve. 

Fig.  47  shows  simi- 
lar curves. 

The  most  common 

condition  is  that  of  constant  speed,  illustrated  in  Figs.  44  and  45.     These 

show  that  at  1150  revolutions  the  head  can  vary  between  0  and  380  feet, 

including  all  friction  losses  and 
suction  lift.  The  curve  indi- 
cated that  the  largest  capacity 
under  no  head  is  2225  gallons, 
and  under  a  head  of  380  feet 
800  gallons,  the  maximum  effi- 
ciency of  70  per  cent  being 
reached  with  a  capacity  of  1200 
gallons,  348  feet  head.  The 
brake-horse-power  curve  shows 
the  amount  of  power  required 
to  run  the  pump  at  constant 
speed  with  a  varying  capacity 
and  head,  the  head  varying 
directly  as  the  square  of  the 
capacity.  The  head  curve 
shows  that  with  the  discharge 
Ive  closed  there  is  a  head  of  285  feet  with  no  discharge.  The  capacity 

increases  as  the  valve  is  opened,  and  the  head  rises  until  it  reaches  380  feet 

for  800  gallons.     At  another  point  where  the  head  is  285  feet  a  capacity  of 

1500  gallons  is  obtained,  giving  the  two  limits  of  operation  of  the  pump. 


800    1000    1200   1400    1600    1800    2004 
Gallons  per  Minute 

Constant  Speed. 


Fig.  44.     Characteristics  Showing  Relations  for 
Constant  Speed. 


50 


CENTRIFUGAL  PUMPING  MACHINERY 


At  constant  speed  and  360  feet  head,  it  would  give  from  400  gallons  to 

1150  gallons. 

^The  following  conditions  should  be  noted: 

First,  when  capacity  Q  equal  to  0  ; 
•^Second,  when  head  H  equal  to  maximum; 

Third,  when  head  H  equal  to  0. 


340 
320 
300 

..280 

& 
£260 

|240 
220 

200 
180 
160 

^~ 

Bead 

j» 

* 

^ 

,s 

70 

\ 

/* 

6°f 

50  §"- 

~4 

^ 

" 

\ 

^ 

< 

e?0" 
_^ 

t^ 

S 

\ 

0 

40  E 

/ 

«<* 

l^ 

N 

V 

\ 

H 

30 

^ 

^ 

\ 

\ 

/ 

\ 

,/ 

\ 

I 

\ 

1  i  1  I 


SII  « 

Gallons  per  Minute 

Constant  Speed. 
Fig.  45. 


I  I  !  I 


Equation  of  Impeller: 

Av2  -  Bv2v  -  cu£  =-2gh. 

w2  =  circumferential  speed,  feet  per 

second. 

v  =  radial  speed  of  water  on  out- 
let, feet  per  second. 
h  =head,  in  feet. 

Example: 

Diameter  impeller  21yf  inches. 
h  =  300  feet. 

n  =  900  rev.  per  unit  (w  =  86  feet 
per  second). 

A=  12.17 

B  =  1.11      for  1  impeller. 
C  =  0.98 
12.17  v2-!. 11  w- 


Curves  A  for  1200  r.p.m. 
Curves  B  for  1000  r.p.m. 
Curves  C  for  800  r.p.m. 
Curves  D  for  600  r.p.m. 

Equation  of  Cap.-Head 

Curve: 
a-Qt-p-n-Q-v-n* 

=  -2gh. 
*  =  17455.4. 
ft  =  19.4335. 
7  =  0.016395625. 

Equation  of  Cap.-Power 
Curve  : 


+  t  -  n\ 

r  =  -  108.6466. 
s  =  0.3407783. 
I  =  0.00008699368. 

Q  =  cu.  ft.  per  sec. 
L  =  sec.  foot  Ibs. 
n  =  r.p.m. 


r 

a 


WV- 


5SA 


\D 


100  200  300  400  500  600  700  800  900  0  100  200  300  400  500  600  700  800  900 
Gallons  per  Minute 

.     Variable, Speed. 

Fig.  46. 


The  maximum  head  does  not  occur  when  the  discharge  valve  is  closed, 
but  at  a  definite  capacity  differing  somewhat  in  the  various  types  of  pumps, 
but  greater  in  high-head  than  in  low-head  pumps,  as  the  volumetric  losses 
in  high-head  pumps  are  greater  than  the  hydraulic  losses.  This  is  illus- 
trated by  considering  the  pump  in  question  as  delivering  into  a  vertical 
pipe  higher  than  the  lift  of  380  feet.  There  would  be  no  discharge,  and  it 


SECOND   ANALYSIS  OR    THEORY 


51 


would  be  impossible  to  start  the  column  again  until  the  head  was  reduced 
to  less  than  285  feet,  or  the  head  which  the  pump  would  produce  with  the 
discharge  closed.  The  maximum  capacity  is  at  zero  head.  In  order  to 

Equation  of  Capacity-Head  Curve: 
a  •  Q2  -  0  •  n  •  Q  -  y  •  n2  =  -  2  gh. 
1  a  =  323.9365. 

TO  0  =  1.6176629. 

7  =  0.005921232. 
60  For  580  r.p.m.: 

h  =  -  5.036  Qz  +  14.587  Q  +  31. 
n  =  head  in  feet. 
Q  =  cubic  feet. 
Equation  of  Capacity-Power  Curve: 

L  =  r  -  n  •  Q2  +  s  •  n*Q  +  t  •  n\ 
*>          r  =  -  0.84913453. 
s  =  0.010273562, 
t  =  0.00001747915. 

•  For  580  r.p.m.: 

*  H.P.  =  -  0.89545  Q2  +  6.2837  Q  +  6.2. 

L  =  sec.  ft.  pound. 
H.P.  =  horse  power. 


800   1000  1200   1400  1600 
Gallons  per  Minute 

Constant  Speed. 


Fig.  47. 


increase  this  capacity  the  speed  must  be  increased,  as  the  capacity  varies 
directly  with  the  speed  at  constant  head.  With  a  constant  head,  and  a 
variable  speed  and  capacity,  the  efficiency  and  horse-power  curves  become 


Capacity 

Fig.  48.  Fig.  49. 

Power  Curves  Showing  the  Effect  of  Different  Vane  Angles. 

different,  and  a  point  is  reached  when  no  increase  in  capacity  can  be 
obtained.     The  power  curve  is 

P  =  rnQ2  +  sn*Q  +  tn*. 


52  CENTRIFUGAL  PUMPING  MACHINERY 

The  horse  power  is  directly  proportional  to  the  square  of  capacity  if  the 
speed  is  constant,  and  directly  proportional  to  the  cube  of  revolutions  if  the 
capacity  is  constant.  When  the  discharge  valve  is  closed  and  there  is  no 
delivery,  the  horse  power  is  directly  proportional  to  the  cube  of  the  revo- 
lutions. With  the  speed  constant,  the  equation  becomes  a  parabola,  with 
the  capacities  as  abscissae  and  the  power  as  ordinates. 

Figs.  48  and  49  show  that  the  variation  in  the  power  depends  upon  the 
angles.  There  are  three  different  parabolas,  1,  2,  or  3,  depending  upon  the 
angles.  We  have  case  1  if  we  select  (az  +  a3)  <  90  degrees;  case  2,  nearly  a 
straight  line,  if  the  angles  (az  +  «s)  =  90  degrees;  and  case  3  if  the  angles 
(az  +  «3)  >  90  degrees.  The  equation  will  solve  the  question  of  maximum 
power  at  constant  speed  for  the  maximum  capacity,  as  Qma*H  =  0. 

APPLICATION  OF  ANALYSIS  TO  PROBLEM. 

The  following  will  illustrate  the  analysis  of  any  problem. 

77    =  total  efficiency  of  pump; 

7)m=  mechanical  efficiency; 

r)v  =  volumetric  efficiency; 

rjh  =  total  hydraulic  efficiency; 

77/11=  absolute  hydraulic  efficiency; 

rje  =  efficiency  motor; 

f]0  =  over-all  efficiency  or  wire  to  water. 

Total  efficiency  is  the  relation  between  the  brake  horse  power  of  motor 
and  the  actual  water  horse  power. 

rr.  4.  i    /c  •  f  water  horse  power      Q  X  H  X  8.33  X  60 

Total  efficiency  of  pump,  r — = — =—  -  =  -  — 

brake  horse  power  33,000 

V  A 

™     ,  .    ,  r      j-  voltage  X  amperes 

Electric  horse  power  for  direct  current  =  —  _._          —  • 

74o 

VXA 

Brake  horse  power,  fje  X     _.      • 

n    rr.  .  water  horse  power 

r/o,  over-all  efficiency,  -. — — . — r—  —  • 

electric  horse  power 

Break-horse-power  output  for  alternating-current  motors 

_  volts  X  amperes  X  Vn  X  cos  <f>  X  77 
746 

where  n  =  number  of  phases; 

cos  </>  =  power  factor  of  motor; 
77  =  motor  efficiency. 

The  speed  of  motors  at  full  load  will  vary  from  2  to  5  per  cent  on  a  motor 
of  200  to  10  horse  power,  and  from  1  to  2  per  cent  from  500  to  200  horse 


SECOND  ANALYSIS   OR    THEORY  53 

power.  Small  motors  of  1  to  10  horse  power  may  vary  from  10  to  15  per 
cent.  The  ratio  between  the  number  of  poles  and  the  speeds  and  cycles 
is  as  follows:  When/  =  frequency,  p  =  number  of  poles.  r  =  number 

of  revolutions;  r  =  120--     The  slip  is  expressed  in  a  percentage  of  the 

actual  revolutions.  Actual  speed  =  revolutions  (1  —  per  cent  of  slip)  .  This 
slip  is  due  to  the  resistance  opposed  to  the  rotor  current. 

The  elements  of  total  pump  efficiency  are,  therefore,  made  up  of 

Mechanical  efficiency  =  rjm. 
Volumetric  efficiency  =  t\v. 
Absolute  hydraulic  efficiency  =  T//^. 

Total  hydraulic  efficiency  =  rjh  =  T?A<  X  Tjr. 
Total  pump  efficiency  =  77™  X  ijv  X  r]ht- 

The  elements  of  mechanical  efficiency  are  made  up  of  friction  of  shaft 
and  impeller,  etc.,  in  their  bearings  and  is  a  function  of  the  workmanship 
and  fits.  The  clearances  vary  from  0.002  to  0.005  inch,  and  should  never 
be  larger  than  0.006  inch.  The  volumetric  efficiency,  rjv,  expresses  the 
ratio  of  the  amount  of  water  entering  the  pump  to  that  which  is  discharged, 
the  loss  being  due  to  leakage  in  the  running  fits.  The  greatest  losses  in  a 
turbine  pump  occur  between  the  impeller  and  the  diffusion  ring.  These 
vary  between  2  to  10  per  cent.  The  absolute  hydraulic  efficiency  riht 
expresses  the  ratio  between  the  useful  head  and  the  total  pumping  head. 
.If  h  represents  the  former  and  H  the  latter,  then  the  total  head  H  is  made 
up  of  h  and  all  frictional  and  shock  losses  of  water  in  the  pump,  and  is  the 
most  important  factor  to  consider  in  designing  pumps.  Information  re- 
lating to  these  losses  is  meager  and  incomplete.  Another  serious  loss  is 
due  to  skin  friction  between  the  impeller  and  the  water.  The  skin  friction 
increases  with  head  pumped  against  more  rapidly  than  the  head  increases 
with  respect  to  the  velocity.  The  head  against  which  the  pump  operates 
varies  as  the  square  of  the  velocity  and  the  wasted  power  as  the  cube  of  the 
head,  or  H3.  The  work  lost  in  disk  or  skin  friction  varies  as  the  square  of 
the  radius,  therefore  a  small  impeller  at  a  high  number  of  revolutions  will 
waste  less  power  than  a  large  one  running  slower,  both  having  the  same 
peripheral  velocity.  The  largest  losses  by  surface  friction  occur  along  the 
walls  of  the  casings,  and  if  the  surfaces  of  the  stationary  walls  and  im- 
pellers were  alike  the  water  would  have  a  rotating  motion  at  a  speed  one- 
half  that  of  the  impeller.  This  loss  may  be  reduced  by  having  the  walls 
smooth  and  the  impellers  polished  and  by  having  proper  clearance  between 
impellers  and  casings.  Experiments  have  been  made  abroad  on  the  power 
loss  by  skin  friction,  and  the  following  formula  has  been  obtained  : 


n* 


54  CENTRIFUGAL  PUMPING  MACHINERY 

W  =  power  due  to  resistance  of  rotating  disk  in  foot  Ibs. ; 

F  =  constant  8132; 

n  =  revolutions  per  minute; 

h  =  head  in  feet. 

In  addition  to  the  surface  friction  of  the  water,  there  are  losses  due  to 
molecular  friction.  The  head  is  shown  to  be  directly  proportional  to  the 
square  of  the  revolutions  and  the  power  lost  on  account  of  friction  to  some 
power  of  revolutions.  The  results  show  that  these  losses  are  greater  in] 
high-head  pumps  than  in  low-head  pumps. 

In  designing  we  may  expect  a  total  efficiency  of  90  per  cent  in  the  large 
pumps,  under  favorable  conditions,  with  a  lower  efficiency  in  the  smaller 
ones.  It  is  absolutely  necessary  to  select  the  right  efficiency  when  den 
signing  this  class  of  pumps,  and  this  can  be  obtained  from  tests. 

The  curve  of  relationship  between  the  capacity  and  efficiency  is  a 
parabola  commencing  at  zero,  its  vertex  showing  the  maximum  efficiency, 
and  coming  down  again  to  zero  as  the  head  approaches  zero.  The  maxi- 
mum point  or  vertex  should  occur  under  the  conditions  for  which  the  pump 
is  designed. 

The  general  equation  for  the  impeller  is 

Av2  -  Bu2v  -  Cu22  =  -2gh, 
where 

u2  =  circumferential  speed  of  impeller; 

v  =  radial  speed  of  water  on  outer  circumference  of  impeller; 
g  =  acceleration; 
h  =  head. 

A,  B,  C  are  constants. 
This  equation  is  analogous  to  the  capacity-head  curve,  v  being  directly 
proportional  to  the  capacity 

FXv  =  Q. 

Q  =  capacity;  F  =  sectional  area. 
u2  is  directly  proportional  to  the  speed 

2riru 
U=^' 

The  general  equation  for  capacity  head  is  a  hyperbolic  parabola, 

aQ2  -(3uQ  -yu2  =  -2gh, 
where  Q,  h,  and  u  are  variables. 

If  n  is  constant  it  becomes  a  parabola  as  illustrated  in  Fig.  46  on  page  50. 
The  figure  shows  the  hyperbolic  paraboloids  for  speeds  at  600,  800,  1000, 
and  1200  revolutions.  Fig.  50  illustrates  the  curve  with  a  constant  head 
and  the  capacity  and  speed  as  variables.  It  represents  an  actual  perform- 
ance under  test.  The  equation  aQ2  —  (3uQ  —  yu2  —  —  2  gh  becomes  a  hyper- 
bola. The  figure  shows  the  relations  between  these  hyperbola  and  the 


SECOND  ANALYSIS   OR    THEORY 


55 


capacity  and  speed  at  constant  heads  of  20,  25,  35,  and  45  feet.  Fig.  51 
shows  the  same  condition  of  a  larger  pump  operated  under  constant  head 
and  a  variable  speed  and  capacity.  . 


650 
Callous  per  Minute 


Fig.  50.     Curves  of  Relation  between  Capacity  and  Efficiency. 

As  an  example,  Fig.  52  will  illustrate  the  method  used  in  designing  im- 
pellers. 

General  equation, 

Av2  -  Buv  -  Cu2  =-2gH, 


3 

= 

^ 

' 

at 

M 

M 

M 

m 

a 

_^—  - 

^ 

>* 

L 

—  —  —  • 

== 

• 

—  - 

— 

I 

i 

833813832 
s     3     3      °°     J§     2     3     ®     2 

Curve  is  Figured                 Gallons 

per  Minute 

Fig.  51.     Curves  of  Relation  between  Capacity  and  Efficiency. 

A,  B,  C  being  constants  or  coefficients.     H  =  head.     2g  =  64.4.     For  v 
and  u  see  Fig.  52. 

r  ,       capacity  in  cubic  feet  per  second  . 

v  =  feet  per  second  =  77^-  — -f r^ —          —7 — -  , 

(D2  X  TT  —  s   X  z)b  in  square  feet 

D2in  ft  X  TT  X  revolutions  per  minute 
u  =  feet  per  second  = 

The  equation 


60 


Av2  -Buv  -  Cv?  =  -2 gH 


56 


CENTRIFUGAL  PUMPING  MACHINERY 


represents  a  hyperbolic  paraboloid,  and  by  taking  u,  v,  or  H  as  constants  the 
three  characteristic  curves  of  the  pump  can  be  obtained,  namely:  Revo- 
lutions constant,  capacity  and  head  varying;  H  constant,  capacity  and 
revolutions  varying;  capacity  constant,  head  and  revolutions  varying. 


Fig.  52.     Double  Suction  Impeller. 

A,  B,  C  depend  upon  the  design  of  pump  and  impeller,  and  can  be 
found  from  previous  available  tests,  or  approximately  from  the  design  of 
the  impeller.  The  impeller  shown  in  Fig.  52,  for  a  66-inch  pump,  gives  the| 
following  values  for  A  =  4.15;  B  =  —  0.17;  C  —  0.98,  making  the  equa- 
tion read 

4.15  v*  +  0.17  uv  -  0.98  u2  =  -64.4#. 


Capacity 
H     Constant 


/'Rev.  per  Minute 
Capacity 
Constant 


Fig.  53.     Equation  Curves  for  Impellers. 


In  order  to  determine  these  constants  from  the  fundamental  equation, 
which  contains  the  variables  Q,  the  capacity;  u,  the  revolutions;  and  H,  the 
head,  one  can  be  taken  as  constant  and  we  can  obtain  three  curves  as 
per  Fig.  53.  These  are  the  characteristic  curves  for  constant  head,  con- 


SECOND   ANALYSIS   OR    THEORY 


57 


slant  speed,  and  constant  capacity. 
The  values  of  A,  B,  and  C  are  usual- 
ly, however,  determined  by  tests, 
for  which  the  following  readings  are 
required  as  per  Fig.  54.  Points  1, 
2,  3  are  test  readings  with  respective 
>Capacity  heads  Hi,  H2)  H3,  and  capacities 
Qi,  $2,  and  Q3.  Introducing  these 

Fig.  54.    Readings  to  Determine  Equation  values  into  the  equation  given,  we 
Curves.  have 


: 

. 
$ 

i 

H, 

1 

H2 

H3 

-Q  » 

Q,       > 

r- 

^3 

AQl-BQu  -Cuz  = 

AQ2  -  BQ2u  -  Cv?  =  -2  gH2, 

AQ3  -  BQ3u  -  Cu*  =  -2gH,. 

The  unknown  constants  A,  B,  C  can  then  be 
obtained. 

The  values  can  be  determined  without  the 
aid  of  tests,  but  it  is  necessary  to  substitute 
in  place  of  capacity  Q  the  relative  velocity  wz 


^          Capacity  ^ 

Fig.  57.     Method  for  Finding  Characteristics. 

of  water  leaving  impeller,  and  in  place  of  u, 
the  revolutions,  the  circumferential  speed  V6 
or  u2,  making  the  equation  as  follows : 
A  -  wz2  -  Bw2u2  -  Cu22  =  -2gH. 

and  uz  are  in  direct  proportion  to  Q  and 
therefore,  referring  to  Figs.  68  and  69, 


Fig.  55.     Various  Forms  of  Im- 
pellers for  Different  Commercial 


68 


CENTRIFUGAL  PUMPING  MACHINERY 


^^  —  =»-          Capacity 

Fig.  56.     Various  Forms  of  Impellers  for  Different  Commercial  Problems. 

f  =  coefficient  of  friction  along  vanes,  approximately  0.1; 
0  =  coefficient  of  losses  due  to  shocks,  approximately  1.2; 

' 


,.     . 
discharge 


Capacity 


velocity;  Fv  =  discharge  area; 
77  =  hydraulic  efficiency. 

SHORTER   METHOD   OF  FINDING 
CHARACTERISTICS. 

A  shorter  method  of  finding  the  char- 
acteristics is  as  follows :  Take  Hf  as  the 
head  with  the  discharge  valve  closed, 

which  by  previous  equation  is  Hr  =  ju—  ; 

fj,  =  varying  between  0.9  to  1.1,  accord- 
ing to  the  proportions  of  impeller.  The 
ordinary  equation  for  a  centrifugal  im- 
peller is 

I 


=  V  - 
r  1 


The  direction  of  the  lines  changes  ac- 
cording to  angle  ai  (see  Figs.  55  and  56)  . 
Assuming  that  we  intend  to  figure  an 
impeller  for  condition  Z  given  in  Fig.  57, 
and  knowing  the  head  H\,  we  can  plot 
capacity  the  curve  from  the  three  elements,  two 

Power  Characteristics  for  Dif-  points  and  a  tangent.     Referring  to  Figs. 

55,  56,  and  58,  the  various  forms  of  im- 


Fig.  58. 

ferent  Shaped  Vanes  of  Impellers. 


peller  vanes  show  the  different  characteristics  found  in  the  usual  type  of 
commercial  pump.    In  No.  1  the  power  increases  considerably,  over  the 


SECOND  ANALYSIS  OR  THEORY 


59 


limit,  and  this  form  is  limited  to  conditions  where  the  load  is  variable, 
due  to  f fictional  resistances  other  than  the  regular  pumping  head. 
No.  2  will  produce  a  constant  head  for  best  efficiency,  but  has  the  disad- 
vantage that  the  power  increases  rapidly.  No.  3,  with  a  constant  speed, 
will  maintain  a  uniform  efficiency  under  a  varying  head. 

The  extent  to  which  a  designer  can  go  in  the  development  of  the  shape  of 
the  vanes  depends  upon  the  conditions  of  service,  and  should  be  closely 
investigated. 

Fig.  59  shows  the  curves  described. 


Curves  A.  Connecting  Points 
of  Highest  Efficiencies  Speed  or 
Revolution  Constant. 

Curves  B.  Connecting  Points 
by  Highest  Efficiencies  Capacity 
Constant. 

Curves  C.  Connecting  Points 
of  Equal  Efficiencies. 

Curves  D.  Showing  Head  for 
Constant  Revolutions. 


Capacity 

Fig.  59.    Characteristics  for  Curves. 


CHAPTER  XI. 


GRAPHICAL  ILLUSTRATION  FOR  DETERMINING  THE 
IMPORTANT   ANGLES. 

IT  is  advisable  to  recall  the  relations  existing  in  plane  trigonometry,  and 
the  measuring  of  angles  in  degrees  and  minutes  in  circular  measure.  Lines 
having  the  names  sines,  cosines,  tangents,  and  cotangents  bear  a  fixed 
relation  to  each  other  for  any  given  angle  and  radius.  Fig.  60  gives  these 
ratios. 


Relation  of  Angles. 

a    .      r,      b  hence  a 
sine  A  =  -  sine  B  =  - 
c  c 


b  „      a 

cos  A  =  -   cos  B  =  - 
c  c 

b 

a 


tang  A  =  7  tang  B 


cot  A  =  -  cotg  B  =  ^ 


c  sin  A 

b  =  c  sin  5 

a  =  c  cos  5 
a  =  6  tang  A 
6  =  acotgA 
6  =  atangJ3 
a  =  6  cotg  B 


Therefore 

sin  A  =  cos  B 

sin  B  =  cos  A 

tang  A  =  cot  B 

tang  B  =  cot  A 


a  =  is  the  sine 
6  =  is  the  cosine 
/  =  is  the  tangent 
m  =  is  the  cotangent 


The  general  equation  as  obtained  in  the  first  part  of  the  analysis  gives 


y,=  V/^-(l- 

f\h 


or 


or 


tang  ao/ 
1 


tang 


cotg 

COtg  (XQ 


cotg /30        cotg 


=  tang,  Fig.  61, 
=  1_  Z5^No>  L 


This  gives  an  equation  for  finding  angle  CKO.  It  can  also  be  found  by 
taking  the  angle  0  between  the  vertical  line  and  the  absolute  velocity  of 
water  (see  Fig.  62)  and  «i,  the  angle  between  the  vertical  line  and  the  rela- 
tive velocity  of  water.  Then 


tang 


tangjS 


(l  - 


and  tang  /3  =  =F  tang  a.\ 
60 


ILLUSTRATION  FOR  DETERMINING  ANGLES  61 


Fig.  61.     Diagram  for  Angles. 


Fig.  62.     Diagram  for  Angles. 


62 


CENTRIFUGAL  PUMPING  MACHINERY 


cti  =  angle  between  radial  and  relative  velocity. 
az  =  90°-  ai. 

ao  =  ai  +  ft  +  flo  =  angle  between  tangential  and  relative  velocity. 
j8  =  angle  between  radial  and  absolute  velocity. 
0o  =  angle  between  tangential  and  absolute  velocity. 


V  =  velocity  of  water  in  discharge. 
Vi  =  radial  velocity  water  at  outlet  in  wheel. 
F2  =  radial  velocity  water  at  inlet  in  wheel  or  Si. 
Fs  =  tangential  velocity  also  used  as  Uz  at  outlet. 
V6  =  tangential  velocity  also  used  as  HI  at  inlet. 
F7  =  velocity  of  whirl  at  inlet. 
Vs  =  velocity  of  whirl  at  outlet. 
F9  =  relative  inlet  speed  also  Wi. 
Vio  =  tangential  velocity  of  water. 
S  =  absolute  outlet  speed. 

Si  =  absolute  inlet  speed  considered  also  as  F2  to  prevent  complication, 
assuming  water  enters  radially  at  entrance  without  shock. 

The  negative  sign  is  used  for  a\  and  0  when  they  are  on  different  sides 
of  center  line,  and  the  plus  sign  when  on  same  side  (see  Fig.  64) . 

Referring  again  to  Fig.  62,  having  found  the  angle  a0,  it  remains  to  find 
angle 

if 

de  =  absolute  speed  of  water 

leaving  impeller; 
ef  =  relative  speed  of  water 


de 


leaving  impeller. 
ef  V, 


sin  a0      sin  0      sin  (a0  — 180) 
(see  Fig.  65). 

Vi  =  radial  speed  =  de  sin  /3o, 

and  by  this  we  obtain 
T,      Tr        sin  a0 


sin  (an  — 


X  sin  /30, 


\ 


or  cotg 


cotg  j30  —  cotgao' 
—  cotg  «o  =  TT  No.  2. 


Fig.  63.     Diagram  for  Angles. 


This  equation,  together  with 
the  one  for  a0,  will  solve  the 
angle  /30-  For  simplicity  in  solving  Fig.  63,  it  is  assumed  that  the  water 
enters  the  impellers  without  loss  and  radially,  hence  F2  is  equal  to  s\. 

The  water  entering  the  impeller  has  a  radial  velocity  72  and  inner  part 
of  the  impeller  a  circumferential  velocity  of  VG.     Let  the  outer  part  of  the 


ILLUSTRATION   FOR   DETERMINING  ANGLES 


63 


impeller  have  a  circumferential 
velocity  F5,  and  let  the  velocity 
of  whirl  at  entrance  be  Vi  and 
at  outer  circumference  F8.  Plac- 
ing Vz  radially  and  ¥7  tangenti- 
ally  a  parallelogram  is  obtained, 
ab  denoting  the  velocity  at 
entrance,  and  by  making  ac 
equal  to  the  tangential  velocity, 
we  obtain  be,  the  velocity  of  the 
wrater  relative  to  the  impeller. 
This  relative  velocity  determines 
the  entrance  angle,  as  the  vane 
must  be  tangent  to  it.  For  the 
outer  diameter,  making  the  velo- 
city of  whirl  Fs,  the  velocity  of 
wheel  F5,  and  the  radial  velocity 
Fi,  the  outer  relative  velocity  ef 
can  be  ascertained,  which  deter- 
mines the  direction  of  the  vane  at 
outer  circumference.  The  work 
done  by  the  water  through  the 

impeller  is  -  (F8  X  F5  -  V7  X  F.) 

y 
foot  pounds  per  pound  of  water, 


—  Minus 


Figs.  64  and  65.     Velocity  Diagram. 

or  assuming  a  radial  entrance 

.      ,              Vs  X  F5 
the  expression  becomes 

Assuming  water  on  entering 
is  turned  from  a  radial  direction 
into  one  of  rotation,  that  its 
velocity  is  F2  and  the  direction 
of  vanes  is  Vg,  the  water  would 
enter  without  shock.  After 
entering  the  actual  velocity  be- 
comes ab.  The  change  of  velo- 
city relative  to  impeller  is  from 
ok  to  ag,  and  the  loss  of  head  is 
(ag  X  aA-)2  (F2  X  ok  -  Vzag) 


Fig.  66.     Graphical  Diagram  for  Angles. 


2g 

=  Fe  —  Vz  cotg 
at  entrance. 


2g 
=  loss  of  head 


Calling  this 


L  =    -  (F6  -  F2  cotg  c^)2  (see  Fig.  63). 


64 


CENTRIFUGAL  PUMPING  MACHINERY 


GRAPHICAL    CHART    FOR    QUICK    REFERENCE. 

This  can  be  laid  out  graphically  (see  Fig.  66),  by  using  the  abscissas  for 
tang  «i  and  tang  /3  as  ordinates. 

Equation  No.  1  will  give  straight  lines,  beginning  at  0,  for  different  values 
of  the  tangential  velocity  F5,  the  efficiency  77^,  and  the  total  head  H. 

Equation  No.  2  will  give  parallel  lines  when  less  than  an  angle  of  45 
degrees  with  the  coordinate  axis. 


Tang.  OL  =Diff vision  Vane  Angle 

5.5  5  4<5  4  3>5 


2.5 


1.5 


•1.4 


7.5 


2.5 


1.5 


This  Line  divided  into  Equal  Parts  advancing  by  -jo 
Fig.  67.     Graphical  Chart  for  Angles. 

The  method  to  be  adopted  is  to  choose  a  tangential  velocity  according  to 
the  speed  and  to  select  a  convenient  diameter  for  the  impeller,  thus  obtain- 

Vi 
ing  a  value  for  1  — 


gh 


which  will  give  one  of  the  lines  radiating  from  0. 

Furthermore,  we  have  the  radial  velocity  V\  =  — _  !r         ; 

TT  •  2  n,  •  W  i 

Wi  being  width  of  impeller  at  the  outside,  Q  being  capacity  in  cubic  feet 
or  gallons  per  second,  R  =  radius  of  the  outside  of  impeller.  When  we  have 

T7 

obtained  values  for  -^  we  will  have  one  of  the  parallel  lines.  The  intersec- 
tion between  the  radiating  line  and  45-degree  line  gives  the  two  angles  a\ 
and  |8  as  shown  in  Fig.  66. 

A  chart  can  be  made  for  ready  reference  in  order  to  determine  graphi- 
cally the  diffusion  vane  and  impeller  angles  as  shown  in  Fig.  67. 


ILLUSTRATION  FOR  DETERMINING  ANGLES 


65 


Vi  V 2 

Parallel  lines  with  a  value  of  ^  and  intersecting  lines  with  a  value  -~ 

K5  0/1 

will  give  tang  a  by  projecting  parallel  to  A-B,  for  a  by  projecting  parallel 

to  C-D. 

>j  EXAMPLE.     See  equation  No.  1. 

Vfa      3504  X  0.80 
0ff-        32.2X64 
where  head  64  feet  =H\  F5  =  58  feet;  ijh  =  0.80;  FI  =  65  feet. 


F5      58 
If  the  diagram  were  laid  out  the  line  1.36  would  intersect  the  ^r  une 

0.112,  and  the  tangent  of  a  would  be  6.65  and  of  «i  2.4.     Hence  a  =  81.5° 
and  ai  67.5°. 

Angle  «3  =  90°  -  67.5°  =  22.5°  (see  Fig.  68),  and  for  the  diffusion  ring 
the  angle  would  be  90°  -  81.5°,  or  8.5°. 

METHOD    OF    CORRECTING    IMPELLER-VANE    ANGLE. 

Assuming  a  pump  designed  for  a  head  of  170  feet  which  on  being  tested 
gave  an  actual  head  obtained  on  test  of  157  feet,  how  much  must  the  vane 
angle  be  increased  in  order  to  get  the  required  head?  (See  Fig.  68.) 


Fig.  68.     Correcting  Angle  at  Tip  of  Vane. 

tang  aix  being  the  new  angle  =  {H  (tang  a  +  tang  «i)  —  tang  a. 
Assuming  a  =  74°  and  «i  =  63°, 

HJ  (3.48  +  1.96)  -  3.48  =  5.02  -  3.48  =  1.54. 

tang  aix  =  1.54  or  57°,  or  it  must  be  raised  the  difference  between  63°  and 
57°,  or  6°. 


CHAPTER  XII. 

THIRD  ANALYSIS  OR  THEORY. 

THEORY  OF  IMPELLERS. 

H  =  head  in  feet. 

Q  =  capacity  in  gallons  or  cubic  feet  per  second. 
n  =  number  of  revolutions  per  minute. 
2  R  =  outside  diameter  in  feet. 
2  r  =  inside  diameter  in  feet. 

F5  =  tangential  velocity  at  outer  circumference  in  feet. 
V6  =  tangential  velocity  at  inner  circumference  in  feet. 
77^  =  absolute  hydraulic  efficiency  or  the  ratio  between  the  useful  head 

h  and  the  total  pumping  head  H. 
g  =  acceleration,  or  32.2. 

a  =  diffusion- vane  angle  at  outer  circumference  of  impellers. 
ai  =  impeller- vane  angle  at  outer  circumference  of  impellers. 
«2  =  impeller- vane  angle  at  inner  circumference  of  impellers. 
W  =  width  of  impeller  at  hub. 
Wi  =  width  of  impeller  at  circumference. 
A  =  area  of  discharge  pipe  in  square  feet. 
AI  =  area  of  impeller  at  outer  circumference  in  square  feet. 

AI  =  2  #  X  TT  X  TFi. 
A  2  =  area  of  impeller  at  inner  circumference  in  square  feet. 

A2  =  2r  X  TV  X  W. 
A3  =  area  at  hub  of  impeller,  the  latter  being  C  in  diameter. 

A3  =  2r2^  -  C2^in  square  feet. 

A  4  =  area  of  suction  pipe  in  square  feet. 

Z  =  number  of  diffusion  vanes. 

Zi  =  number  of  vanes  in  impeller  at  outer  circumference. 
Zi  —  number  of  vanes  in  impeller  at  inner  circumference. 
It  is  customary  to  make  Zi  =  2  Z2. 

/  =  thickness  of  diffusion  vanes. 

ji  =  thickness  of  impeller  vanes  at  outer  circumference. 

/2  =  thickness  of  impeller  vanes  at  inner  circumference. 

y  =  velocity  of  water  in  discharge  pipe. 
Vi  =  radial  velocity  in  section  AI  of  impeller. 
V2  =•  radial  velocity  in  section  A2  of  impeller. 
V3  =  velocity  in  section  As. 
V^  =  velocity  in  suction  pipe. 

66 


THIRD  ANALYSIS  OR   THEORY  67 

To  prevent  shocks  and  losses,  an  impeller  must  be  so  designed  that  the 
velocities  will  increase  gradually  in  going  from  section  A3  to  A\. 

The  most  important  values  required  are  the  inner  and  outer  vane  angles, 
which  may  be  determined  by  the  following  equations : 


F5  =  V  —  v  (1  -f-  cotg  a  x 


or  5  =  V  —  C1 

r  *?fti 

V 


—r-  --  r, 

g  (1-f  cotgatangai) 

V^h 
cotgatangai  =     -     -  1. 


and  H 


The  angle  a  is  usually  between  70  and  83  degrees,  depending  somewhat 
upon  the  capacity. 

The  absolute  hydraulic  efficiency  to  be  selected  is  between  69  to  80  per 
cent,  depending  upon  whether  the  pump  is  of  the  volute  or  diffusion  type, 
and  also  upon  the  head,  as  it  governs  the  diameter  of  the  impeller. 

The  angle  a  is  such  that  the  velocity  F5  is  larger  than  V2)  but  care  must 
be  taken  that  the  space  W\  does  not,  with  small  angles,  become  too  small. 
When  the  value  of  a  has  been  calculated  we  have 


tang  a 
tang  ai  =  --  --*  -  -  tang  a, 


V,  r 


Example  No.  1.  —  Capacity,  25,000,000  gallons  per  24  hours;  head,  200 
feet.  Dividing  this  head  into  two  stages  of  100  feet  each,  we  find  as  follows: 
Size  of  suction  and  discharge:  Selecting  a  water  speed  in  feet  per  second 
from  6.6  to  8.2  feet,  we  will  find 

2325  2325 

*  =  t0  Square  feet> 


=  4.72  to  5.87  square  feet, 
=  29.4  to  32.8  inches  diameter. 
We  therefore  select  a  30-inch  discharge,  and  as  it  is  desirable  to  have  the 
suction  pipe  a  size  larger  a  36-inch  suction  is  chosen  in  order  to  cut  down 
the  friction  and  other  losses. 

To  find  the  inside  diameter  of  impeller,  or  2  r  :  The  velocity  of  the  water 
entering  the  impeller  at  A3  is  F3  can  be  determined  by  the  empirical 
formula,  73  =  0.09  to  0.12  \/2  gH,  in  which  g  =  32.2  and  H  =  100  feet. 

F3  becomes  7.23  feet  to  9.64  feet  per  second. 

Selecting  9.2  feet  per  second  and  remembering  that  V3  must  be  larger 
than  74,  we  find  that 

cotga  X  tang«!  =  -* 


68  CENTRIFUGAL  PUMPING  MACHINERY 

Taking  tang  a  —  4.2,  we  obtain 

79.12  X  0.83  X  4.2 
tangai=         32.2  X  100          -4-2>or2-8> 

c^  =  tang' 

This  angle  gives  the  radial  velocity  V\  of  11.5  feet  per  second  of  the  water 
at  the  circumference  of  the  impeller,  which  is  satisfactory,  as  it  is  larger 
than  F2  and  F3. 

Width  of  impeller  at  circumference,  Wi,  and  at  hub  =  W. 

In  order  to  determine  this  the  impeller  should  be  laid  out  on  the  drawing 
board  and  from  the  information  already  obtained  the  internal  passages 
can  be  properly  determined,  and  with  the  selection  of  the  number  of  vanes 
the  space  can  be  calculated.  In  laying  the  impeller  out  it  is  well  to  allow 
10  per  cent  for  losses  in  the  passage  through  it.  The  angle  o^  of  the  im- 
peller at  inner  circumference  can  be  obtained  as  follows: 

VK         39  6 

tang  a:2  =  -7-  =  ypr^ ,  or  3.66, 

where  A%  =  10.8  feet  per  second, 

and  Fe  =  39.6  feet  per  second. 

2325 
The  area  at  A3  will  be  •      '       9  =  4.2  square  feet;  adding  0.5  square  foot 

OvJ     /N      t/.^J 

for  the  hub,  it  becomes  4.7  square  feet.     The  diameter  of  impeller  at  the 
inside  is     /  -—  ,  or  2  feet  6  inches  diameter.     This  would  give  a  very  slow 

4 

speed  and  would  require  a  large  wheel,  hence  it  would  be  best  to  increase 
the  velocity  at  the  entrance  of  impeller  to  10.8  feet  per  second,  which  would 

2325 

give  A3  =  ™ iyr^  =  3.6  square  feet,  plus  0.5  square  foot,  or  4.1  square 

uU  X  -LU.o 

feet  in  all,  or  a  diameter  of  2  feet  3J  inches.     It  is  recommended  that  the 
outside  diameter  of  the  impeller  2  R  =  4  r,  which  in  this  case  would  give 
2R  =  4  feet  7  inches  diameter,  or  R  =  2  feet  3J  inches. 
The  circumferential  velocity  of  the  impeller: 

/    1.40. #      c/1.4X  32.2X100     „_, 

This  being  V 5  =     /  -    —775 — ^  =  V i 77T^\2 —  =  • '  •"  ^ee^  Per  second. 

/    -|      f*r\  ~  (v-b) 

4  /        1  — 


The  number  of  revolutions  would  then  become 

60  X  V,      60 .  77.6 
n= jr-fr  =  -     ^-^~  —  325  revolutions  per  minute. 

7T  •  Z  K  7T  •  4. DO 

Calling  this  330  revolutions,  the  speed  F5  becomes  79.1  feet  per  second, 

-17 

and  Vz  =  -£  =  39.6  feet  per  second. 


THIRD  ANALYSIS  OR   THEORY  69 

The  angles  of  vanes  may  be  calculated  as  described  on  pages  63  and  64, 
[Chapter  XI. 

a  =  diffusion- vane  angle; 

«i  =  impeller- vane  angle; 

rjh  =  absolute  hydraulic  efficiency,  assumed  at  83  per  cent. 

The  method  thus  described,  with  the  velocity  diagram,  Chapter  XI,  will 
give  the  principal  details  governing  the  calculation  of  a  turbine  pump. 
I  The  shape  of  the  vanes  must  be  a  continuous  curve.  In  order  to  start  the 
pump  against  a  full  head  the  velocity  of  impeller  F5  must  be  larger  than 


1.35  or  1.4  X  32.2  H 


i-^Y 


APPLICATION    OF    ANALYSIS    TO    PROBLEM. 

Example  No.  2.  —  8-inch  3-stage  turbine  pump. 
Conditions.  —  1800  gallons  or  241  cubic  feet  per  minute; 

250  feet  head,  or  83.3  feet  per  stage; 

800  revolutions  per  minute. 

Volumetric  efficiency,  95  per  cent;  total,  65  per  cent. 
Hydraulic  efficiency,  f  f  =  68  per  cent. 

Diameters  of  impellers:  hub,  2^  inches,  assumed;  inside  diameter,  2 
inches,  assumed;  outside  diameter,  2  R  =  18f  inches,  assumed. 
The  outside  diameter  gives  a  speed  of 


65  feet  per  second. 


uU 

Speed  required  to  start  pump  against  full  head, 


/lAxgXH          /1.4X32.2X83.3 

F5  >        --  =         -  — 


=  64.3  feet  per  second. 


Velocity  of  the  water  in  8-inch  discharge, 


241 
V  =  -        .  \  /  o  \2  =  11-5  feet  per  second. 

60Xm(r7r  " 


Assuming  a  suction  pipe  the  same  diameter  as  the  discharge,  the  speed  with 

95  per  cent  volumetric   efficiency  would  be  F4  =  ^-^  =  12.1   feet   per 

u.yo 

second.  This  is  high,  hence  it  would  be  well  to  select  a  10-inch  suction.  A 
pump  designed  upon  the  basis  of  the  given  piping  would  probably  not  give 
over  65  per  cent  efficiency.  Calculating  the  angles  a  and  «i  would  further- 
more show  that  an  8-inch  diameter  would  not  be  as  suitable  for  these  con- 


70  CENTRIFUGAL  PUMPING  MACHINERY 

ditions,  as  a  10-inch  pump,  and  that  the  speed  should  be  increased  to 
about  1000  revolutions  in  order  to  get  good  results. 

Analyzing  the  angles,  a  and  ai  have  to  be  taken  so  that  they  will  make 
Vi  >  F3.     We  have,  therefore, 

241.  122 
F3  =  -  -  -   =  20.5  feet  per  second. 

60^(6.52-2.52) 
Taking  tang  a  =  3,  we  have  tang  ai  =  tang  a\  —  ^~  —  1 

L  gu.        j 

_  65*  X  0.68 

3X 


32.2X83.3~M 

and  Fi  =  -  -  •  =  =  20.3  feet  per  second. 

tang  a  X  tang  «i       3  +  0.21 

These  angles,  a  =  72°  and  «i  =  12°,  are  too  large  for  good  results. 
Example  No.  3.  —  4-inch  12-stage. 
Conditions.  —  150  gallons  or  20  cubic  feet  per  minute; 
610  feet  head,  or  50.9  feet  per  stage; 
1460  revolutions  per  minute. 
Efficiency,  volumetric  =  90  per  cent. 
Total  efficiency  =  56  per  cent. 

Hydraulic  efficiency      =  62  per  cent. 

Diameter  of  impellers:  hub,  2  inches;  2  r  =  3|  inches  inside;  2  R  =  9 
inches  outside. 

Speed  for  9-inch  impeller  =  F5  =  -  ~^o  --  =57  feet  per  second. 
Speed  to  start  under  full  head, 


T,  .        /  1.4X32.2X50.9      K0  ,  , 

F5  >     /  -          /0  Q>7K>2 —  =  52  feet  per  second. 


/3.875\2 
(    9    / 


Speed  of  water  in  suction  and  discharge,  assuming  the  same  diameters, 

20  1 
F  =  -  -  =  3.83  feet  per  second, 


3  83 

F4  =  7  ;r  =  4.26  feet  per  second. 
u.y 

It  would  be  advisable  to  make  suction  pipe  5  inches  diameter  in  order  to 
reduce  velocity. 

Velocity  F3  =  -  -  =  5.11  feet  per  second. 

60  •     (3.S752  -  22) 


THIRD   ANALYSIS   OR    THEORY  71 

<Take  tang  a  =  8,  which  will  give 


and  Vi  =  Q-        ^7  =  5.8  feet  per  second. 

o  ~\~  I.o4 

Thus  Tri  >  F3,  as  it  should  be. 

Angle  «2,  speed  72  =  5  feet.       2x11    =  5* 

3  875 

speed  F6  =  57  X  —     -  =  24.6  feet  per  second. 
y 

24.6 

tang  «2  =  —  F~  =  4.92. 
o 

Assuming  8  diffusion  vanes  and  12  impeller  vanes  at  outlet,  with  thick- 
ness of  diffusion  vanes  H  mch  and  impeller  vanes  •/?  inch,  we  have 
20  (  /0.6875inch\  /0.28 


12 

0  ^^  V  19 


-  -- 

o.o 


J2.36  -  8  X  0.057  -  12  X  0.023J, 


TFx  =  0.69  -*-  J2.36  -  0.456  -  0.276], 

Wi  =  ¥J>  =  0.42  inch. 
l.oo 

lowing  10  per  cent  for  losses,  the  width  W\  becomes  0.46  inch,  or  about 
TB-  inch. 

Inside  width  W,  therefore,  may  be  similarly  calculated,  assuming  6  vanes 
at  entrance  J  mch  thick. 


W  =     5pr  y      -5-  f  3.14  X  0.32  -  6  X  0.02}, 

ft  7QQ 
W  =  0.799  -HI-  0.12}  =  7     ^  =  0.896  inch. 

U.oo 

Allowing  for  10  per  cent,  W  =  0.985,  or  about  1  mch. 

It  should  be  noted  that  hydraulic  efficiencies  of  62  to  79  per  cent  should 
be  used  for  volute  type  of  pumps,  about  62  to  70  per  cent  for  small  turbine 
pumps,  and  70  to  85  per  cent  for  large  ones,  the  value  varying  with  the 
type  and  design  of  the  particular  pumps  considered. 

A  method  which  may  be  applied  in  a  somewhat  different  manner  is  given 
below. 

Radial  velocity  at  outer  circumference,  V\\ 
Radial  velocity  at  inner  circumference,  F2; 
Wi  =  width  at  outer  circumference; 

W  =  width  at  inner  circumference; 

/i  =  thickness  of  vane  at  outer  circumference; 

/2  =  thickness  of  vane  at  inner  circumference; 

k  =  coefficient  of  contraction  =  0.90. 


72  CENTRIFUGAL  PUMPING  MACHINERY 

Q 


/    • 

=  ej  sin  0:3  = 


-7^.  —  ^^ 
(27TJKTF  !  - 


(2irrW-nWcoseca4f2)k 
See  Fig.  62. 

The  angle  a3  varies  in  practice  from  15  to  30  degrees. 

The  width  W\  and  W  can  be  calculated  from  the  following  formulae: 

Q 


cos  a      cos  ai 

Q 


Z   =  number  of  diffusion  vanes. 

Zi  =  number  of  impeller  vane's  at  outlet. 

Z2  =  number  of  impeller  vanes  at  inlet. 

These  values  of  W  should  be  increased  about  10  per  cent  for  interna 
leakage  and  losses. 

Volumes  should  be  calculated  in  cubic  feet  per  second  and  dimensions 
in  feet. 


CHAPTER  XIII. 
SCREW  OR  PROPELLER  PUMPS. 

THE  latest  development  of  the  centrifugal  pump  is  its  use  combined  as  a 
unit  with  the  steam  turbine  for  circulating  water  at  very  low  heads,  and 
in  moving  large  bodies  of  water  under  low  velocities  and  at  comparatively 
low  heads.  In  this  field,  high  speed  causes  the  designer  special  difficulties 
in  determining  the  proper  inlet  and  outlet  diameters  of  the  impellers,  and 
"the  length  of  the  vanes;  and  frequently  the  path  of  the  water  in  the  impeller 
becomes  too  short.  For  such  work  the  moving  of  the  water  in  an  axial 
direction  is  necessary,  and  a  screw  or  propeller  is  particularly  adapted  to  it. 
Screw  pumps  belong  to  the  same  class  of  velocity  pumps  as  centrifugal. 
They  consist  of  guide  vanes  at  the  inlet,  with  screw  or  propeller  and  guide 
vanes  at  the  outlet.  All  are  mounted  in  a  cylinder  or  chamber.  The 
velocity  of  the  water,  being  parallel  or  axial,  lends  itself  readily  to  operation 
at  high  speeds,  with  high  efficiency  for  water  at  low  heads.  By  arranging 
the  screws  or  propellers  in  opposition  to  one  another,  all  end  or  lateral 
thrust  on  the  shaft  is  eliminated,  and  at  the  same  time  the  capacity  can 
be  doubled  without  increase  of  speed.  The  angles  of  the  stationary  inlet 
and  outlet  vanes  and  blades  of  the  screws  are  so  arranged  that  the  water 
enters  axially  and  is  given  a  radial  motion,  finally  being  discharged  parallel 
to  the  axis  in  the  vortex  or  volute  chamber,  thereby  utilizing  both  the 
impulse  and  reactional  forces. 

In  a  centrifugal  pump  the  liquid  flows  in  a  radial  direction  through  the 
impeller  and  obtains  its  energy  through  the  difference  in  circumferential 
speed  at  the  inlet  and  outlet  of  the  impeller.  In  a  screw-propeller  pump 
the  liquid  flows  axially,  there  is  no  increase  in  circumferential  speed,  hence 
the  liquid  must  obtain  its  pressure  through  some  other  cause. 


Fig.  69.     Path  of  Water. 

Suppose,  first,  that  a  screw  propeller  revolves  in  the  water  so  that  no 
shock  occurs  either  at  the  inlet  or  outlet.  The  water  flows  as  shown  in 
diagrams  at  A  and  B  (see  Fig.  69). 

73 


74 


CENTRIFUGAL  PUMPING  MACHINERY 


u-i,  circumferential  speed  at  A,  is  equal  to  uz,  circumferential  speed  at  B, 

vroJ  absolute  inlet  speed,  equal  to  tva,  absolute  outlet  speed. 

w\}  relative  inlet  speed,  equal  to  Wz,  relative  outlet  speed. 

The  diagrams  at  A  and  at  B  are  similar.  The  liquid,  therefore,  has  the 
same  energy  at  A  and  B.  During  the  flow  of  the  liquid  from  A  to  B  no 
energy  has  been  put  into  the  liquid  and  no  pumping  head  can  be  produced. ' 
A  smooth,  shockless  flow  through  a  screw  propeller  does  not  produce  any  pump- 
ing head. 

Slip.  —  A  certain  shock  must,  therefore,  be  produced  in  order  to  obtain 
pressure.  The  full  theoretical  capacity  cannot  pass  through  the  propeller. 
This  reduction  in  capacity  is  called  slip.  Instead  of  letting  the  water 
enter  as  per  diagram  uv row,  we  let  it  enter  as  per  diagram  ucwf  (see  Fig.  70) . 


1C '    Diagram  with  Slip 


Fig.  70.     Entrance  Angles  for  Water. 

Definition  of  Slip.  — 

Qo  =  capacity  without  slip,  or  the  theoretical  capacity. 
Q  =  capacity  with  slip,  or  the  actual  capacity. 
z  =  slip. 


^~  =  1  —  z  =  ratio  between  obtained  capacity  and  maximum  capacity. 


Section  through 
Propeller 


Development  of  Vanes  at  Diameter  D 


Fig.  71.  Guide  Vanes. 

Assuming  radial  guide  vanes  we  find  as  in  Fig.  71, 
Qo  =  A  •  vro  =  A  -  u  •  tang  5. 


A  =  radial  area  =  [D22 

5  =  pitch  angle. 

z  =  number  of  vanes. 


u  =  circumferential  speed  = 


_  D  »7r  •  r.p.m. 


60 


SCREW  OR   PROPELLER   PUMPS 


75 


Assuming  guide  vanes  entering  at  an  angle  e  in  direction  of  rotation  (see 

Fig.  72), 

Qo'=  theoretical  maximum  capacity; 


tang  e/ 


Fig.  72.     Guide  Vanes. 

Assuming  guide  vanes  entering  at  an  angle  e,  pointing  opposite  to  direc- 
tion of  rotation,  .    . 

Qo"  =  theoretical  maximum  capacity; 


For  Qo'  the  only  difference  in  the  equations  is  in  the  values  of  tang  e,  which 
are  positive  and  negative  respectively  for  $</  and  Q0". 

Qo"  >  Qo'. 

If  the  inclination  of  the  guide  vanes  is  in  the  direction  of  rotation,  the 
maximum  capacity  for  slipless  flow  is  smaller  than  for  radial  guide  vanes. 
For  guide  vanes  pointing  against  the  direction  of  rotation,  the  maximum 
capacity  is  larger  than  for  radial  guide  vanes: 

Qo'  <  Qo  <  Qo". 

2  =  1-^-  fore  =  90°, 


z  =  l- 


90°, 


<  QQ'J  Qo"  being  figured  by  the  above  equations. 

Law  of  Proportionality  and  Slip.  —  For  water  turbines  and  centrifugal 
pumps  a  law  exists  which  can  be  expressed  by  the  formula 


76 


CENTRIFUGAL  PUMPING  MACHINERY 


where  N  =  revolutions  per  minute; 

Q  =  capacity; 
H  =  total  head; 
H.P.  =  horse  power. 

This  law  implies  that  the  capacity  changes  in  direct  proportion  to  the 
speed,  the  head  changes  in  proportion  to  the  square  of  the  speed,  and  the 
horse  power  changes  approximately  as  the  cube  of  the  speed.  This  law 
has  been  found  to  be  correct  within  practical  limits. 

Propeller-pump  tests  show  the  interesting  fact  that  this  law  of  propor- 
tionality can  also  be  applied  to  ship  propellers. 

For  low  speeds  the  tests  show  conformity  in  regard  to  speed,  capacity, 
and  head;  for  the  higher  speeds,  however,  and  for  the  horse  powers,  some 
irregularities  occur.  The  law  of  proportionality  can  be  shown  by  the  curves 


Q 


Q 


-^  =  constant  and  — ^  =  constant.  See  Fig.  73. 


\ 


SlipC 


Q 
LU 


Constant  Slip  Curv 
'for  Maximum  Efficiency 


These  |  Curves  are  made 
from  Test  Headings, 
Two  Propell  rs  working 


Profiler 
Dimen'sior 


0  1000  2000  3000  4000 

Gallons  per  Minute 

Fig.  73.     Performance  Curves. 
It  can  be  proved  that  if  the  law  of  proportionality  is  correct  then  the 

curves  -^  =  constant  and  -4=-  =  constant  represent  curves  of  constant  slip. 
J\  \  H 


If 


^  =  constant,      jj  =  constant  =  law  of  proportionality. 


9. 

N 


—  constant,  then  z  =  constant. 


SCREW   OR   PROPELLER   PUMPS 


77 


This  result  is  of  great  importance,  as  it  shows  that  points  of  same  slip 
must  be  points  of  similar  efficiency. 

For  changes  of  speed  the  best  efficiency  will  occur  at  a  constant  slip. 
For  differently  designed 
propellers  the  best  effi- 
ciency will  occur  at  dif- 
ferent slips.  For  a  pro- 
peller of  dimensions 

given  in  Fig.  73,  the  best  ?4  er  ^  without  Q 

efficiency,  66   per  cent 

occurred  at  1800  r.p.m.  with  37  per  cent  slip.    Fig.  74  shows  the  position 
of  vanes,  with  free  space  or  clearance  and  no  overlap. 

^^--  Characteristic  Curve.  — 
Since  slip  is  necessary  in 
order  to  produce  pres- 


Total  Number 
of  Vanes,  Three 


Fig.  75.     Step  Diagram. 


sure,  the  relation  between  slip 
and  pressure  must  be  found. 
This  theory  is  based  upon  the  assumption  that  all  the  shock  produced 
[>y  the  slip  is  transformed  into  pressure,  and  that  the  pressure  can  be 
igured  from  this  shock, 
provided  a  proper  co- 
efficient is  applied. 

On  this  basis  we  have 
the  formula 


120 
110 
100 
90 

80 

a 
S 

50 
40 
30 
20 

10 

0 

i 

\ 

\ 

•**. 

\ 

7 

~ 

P- 

\ 

\ 

/ 

1 

\ 

/ 

1 

\ 

/ 

\ 

x 

\ 

"**/ 

N^ 

\ 

^^^s 

A=i 
B=c 

68t   CU 

irveflj 

-ve  for 
uredf 

1800  B 

•om  Eq 

P.M 
oatioii 

\ 

\ 

^ 

SN 

\ 

B 

H  = 

*     1? 

-o- 

siuo) 

X 

\ 

,-i 

urve  fi 

T  Coeft^ 

cient  • 

\ 

s^ 

\ 

^ 

\ 

D,- 

6*" 

°2= 

11M" 

5= 

13° 

^ 

^^ 

L. 

>                       1000                     20UU                    3000                     4000                    5000          Q      601 

Gallons  per  Minute                                   xo 

z  =  slip; 

u  =  average   circum- 
ferential speed; 
g  =  acceleration  of    so 

gravity; 
a  =  absolute    angle 

formed   by  the 

direction  of  the 

inflowing  liquid 

with  the  moving 

propeller  vane; 

y  =  coefficient. 

^      T^-      rjr  Fig.  76.     Application  of  Equation  Described. 

While  this  equation  does  not  give  reliable  results,  it  can,  if  properly 
applied,  be  used  to  give  an  approximation.     Fig.   76  shows  how  this 

equation  can  be  used. 

*  Hollander's  formula. 


78 


CENTRIFUGAL  PUMPING  MACHINERY 


The  coefficient  y  stays  nearly  constant  for  a  large  portion  of  the  curve; 
for  the  lower  heads,  however,  it  changes  suddenly.  It  changes  for  various 
propellers  and  working  conditions,  and  its  value  must  therefore  be  deter- 
mined by  tests  in  each  particular  case. 


Fig.  77.     General  Arrangement  of  a  Screw  Pump. 


Fig.  78.     Section  of  a  Screw  Pump. 

A  propeller  pump  is  subject  to  the  same  general  law  of  proportionality 
as  water  turbines  and  centrifugal  pumps.  A  smooth  water  flow  does  not 
produce  pressure :  shock  is  a  necessity.  A  high-speed  propeller  pump  can, 
therefore,  never  have  the  same  efficiency  as  a  properly  designed  high-speed 
centrifugal  pump. 


SCREW   OR   PROPELLER   PUMPS 


79 


The  general  arrangement  of  a  screw  pump  is  shown  in  Fig.  77,  which  gives 
'approximate  outside  dimensions,  and  in  Fig.  78,  showing  the  pump  in 
section.  Fig.  79  is  a  section  showing  the  combination  of  screw  and  volute 
centrifugal  pump,  adapted  for  a  very  low  head.  The  inlet  is  trumpet- 
shaped  to  receive  the  current  of  water  with  the  least  loss.  The  impeller  is 


Fig.  79.     Section  Showing  the  Combination  of  Screw  and  Volute  Centrifugal  Pump. 

j  formed  by  helicoidal  vanes  generated  from  the  lowest  part  of  the  impeller 
]  and  pitched  backwards  like  a  screw.  From  this  form  they  gradually 
:  change  into  the  regular  shape  of  volute  impeller  vanes.  The  discharge 
*  passages  can  form  a  regular  volute  casing  when  water  is  to  be  delivered 
,  into  a  pipe  or  conduit.  For  canal  work  the  open  discharge  casing  is  suit- 
!  able  in  connection  with  the  necessary  sluice  valves  in  the  canal. 


PART   III.— APPLICATIONS   AND   USES. 


CHAPTER  XIV. 

GENERAL  REMARKS. 
WATERWORKS  INSTALLATION. 

OWING  to  its  characteristics  the  centrifugal  pump  is  better  adapted  to 
some  engineering  problems  than  to  others.     Its  simplicity  of  construction, 


Fig.  80.    Vertical,  Self-contained,  Two-million  Gallon  Turbine  Pump. 

wide  passageways,  absence  of  valves,  low  first  cost,  comparatively  light 
weight  for  its  capacity  and  its  adaptability  for  motordrives  have  given  this 
type  of  pump  a  constantly  widening  field  of  application  during  the  last 
fifteen  years. 

81 


82 


CENTRIFUGAL  PUMPING  MACHINERY 


GENERAL   REMARKS 


83 


In  this  and  following  sections  are  described  the  more  general  and  impor- 
tant :r^~l:  cut  ions  of  the  centrifugal  pump  up  to  the  present  time. 

WATERWORKS. 

Turbine  pumps  are  now  supplying  water  to  municipalities  like  Buffalo, 
Lockport,  Toronto,  Montreal,  Louisville,  Minneapolis,  and  others,  and 
furnish  a  constant  and  uninterrupted  supply.  The  sizes  range  from  small 
plants  of  a  million  gallons  per  day  to  those  of  twenty-five  to  thirty  mil- 
lion gallons  per  day.  In  applying  centrifugal  pumps  to  service  of 
this  sort,  the  best  practice  calls  for  delivery  to  the  top  of  standpipes  or 
reservoirs.  In  this  way  both  the  capacity  and  pressure  can  be  kept 
constant. 


Fig.  82.     Twelve-inch  Three-stage  Pump. 

Fig.  80  illustrates  a  vertical,  self-contained,  two-million-gallon  turbine 
pump  installed  at  Athens,  Ga.,  a  type  which  takes  up  little  floor  space,  and 
which  has  proved  very  satisfactory  for  small  waterworks  and  manufactur- 
ing plants. 

Figs.  81  and  82  illustrate  the  three  12-inch  three-stage  pumps  of  the 
waterworks  plant  of  the  city  of  Lockport,  N.  Y.  Fig.  83  is  the  test  curves 
of  the  motor. 

The  pumps  are  directly  connected  to  500-horse-power  motors  and  are  of 
the  horizontal-shaft  three-stage  turbine  type,  with  a  single  suction.  The 
suction  openings  are  14  inches  in  diameter  and  fitted  with  special  vapor 
openings.  The  main  casing  is  cast  iron  of  a  tensile  strength  of  30,000 
pounds,  annular  in  form,  and  fitted  with  suitable  supports  to  attach  to 
base.  The  suction-head  casting  is  of  special  design  to  facilitate  the  re- 
moval of  internal  parts  without  dismantling  the  pumps.  The  impeller  is 
special  bronze,  of  the  inclosed  type,  and  is  arranged  with  balancing  cham- 


84 


CENTRIFUGAL  PUMPING  MACHINERY 


bers,  which  reduce  the  end  thrust  to  a  minimum.  All  impellers  are  per- 
fectly balanced  individually  and  when  mounted  together  on  the  shaft. 
The  discharge  from  each  impeller  is  conducted  through  a  set  of  guides  or 
diffusion  vanes,  designed  to  transform  the  velocity  into  pressure  with  the 
least  possible  loss.  These  diffusion  vanes  are  removable.  The  shafts  are 
of  nickel  steel,  ground  and  polished,  and  run  in  ring  oil  babbitted  bearings. 

Stuffing  boxes  are  fitted  with 
water  seals,  consisting  of  lan- 
tern glands  connected  up  with 
suitable  piping,  obviating  all 
air  leaks.  All  nuts  are  case- 
hardened,  and  the  backs  of 
the  flanges  spot-faced.  The 
installation  is  a  good  sample  of 
a  modern  turbine  waterworks 
installation.  Each  pump  has 
a  capacity  of  5,000,000  gallons 
of  water  in  24  hours,  when 
operating  at  speeds  given  in 
the  test.  The  motors  were 
intended  to  run  at  720  r.p.m., 
but  actually  run  at  745  r.p.m. 
The  pumps  deliver,  therefore, 
considerably  more  than  the 
contract  requirement,  and 
maintain  their  efficiency. 
They  were  guaranteed  to  give 
68  per  cent  efficiency,  but  ac- 
tually give  82.5  per  cent. 

Water  consumption  in  the 
city  of  Lockport  is  between  400,000  and  500,000  gallons  per  day  and  the 
water  is  delivered  through  68,500  feet  of  30-inch  steel  pipes  from  the  Niagara 
River  at  Tonawanda,  into  a  standpipe  of  limited  capacity  near  Lockport. 
The  test  was  made  with  a  constant  level  in  the  standpipe  at  Lockport, 
by  throttling  the  discharge  valves  on  the  pumps  until  the  Venturi  meters 
showed  that  the  pump  was  delivering  at  a  five-million-gallon  rate.  The 
speed  was  745  r.p.m.  The  gate  valves  were  then  opened  and  records  taken 
for  ratings  of  five  and  a  half  million,  six  million,  seven  million,  and  seven 
and  three-quarter  million  gallons. 

The  following  log  of  these  tests  is  of  interest  because  it  shows  £he  extraor- 
dinarily high  efficiency  and  capacity  of  comparatively  small  pumps  for 
waterworks  service.  A  record  was  also  kept  of  the  pressures  in  the  force 
main,  the  rate  of  pumping  resulting,  and  the  horse  power  of  the  motor  as 


500  1000  1500  2000  2000  3000  3500  4000  4000  5000  5500 
Torque 

Fig.  83.     Performance  Curves  of  Motor  Driving 
Pump,  shown  in  Figure  83. 


GENERAL   REMARKS 


85 


indicated  by  the  wattmeter 
on  the  gauge  board.  A  com- 
plete record  of  the  test  is 
shown  in  the  table. 

Repeated  tests  on  turbine 
pumps  after  years  of  serv- 
ice have  shown  that,  when 
rightly  constructed,  there 
is  little  or  no  falling  off 
in  efficiency,  a  condition 
hardly  met  with  in  recipro- 
cating pumps,  where  wear 
of  valves  is  always  present. 
In  considering  a  waterworks 
or  mill  installation,  original 
cost  and  maintenance  should 
both  be  taken  into  account, 
as  the  turbine  pump  has 
many  advantages  over  the 
plunger  pump  in  lubrication, 
repairs,  or  replacement  of 
operating  parts,  together 
with  first  cost,  foundation, 
and  building. 

They  should  be  designed 
for  a  fairly  wide  range  of 
discharge,  and  usually  for 
a  constant  head.  Over-all 
efficiencies  of  from  68  to  72 
per  cent  can  be  obtained  as 
against  75  to  80  per  cent  for 
the  reciprocating  pumping 
engine.  The  latter  still  has 
an  advantage  in  the  cost  of 
power,  which,  however,  is 
offset  by  the  lower  cost  of 
maintenance  of  turbine 
pumps,  of  their  buildings, 
foundations,  interest  on  in- 
vestment, depreciation,  and 
repairs. 


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86  CENTRIFUGAL  PUMPING  MACHINERY 


TESTS  OF  CENTRIFUGAL  PUMPING  ENGINES,  MONTREAL. 

The  results  of  the  tests  of  the  engine-driven  turbine  pump  at  the  low- 
level  station  of  the  Montreal  waterworks  have  recently  been  made  pub- 
lic. The  plant  consists  of  a  Worthington  three-stage  turbine  pump  and 
a  750-horse-power  triple-expansion  steam  engine  as  shown  in  Fig.  83A. 
The  contract  required  that  the  unit  should  have  a  duty  of  100,000,000 
foot-pounds  for  each  1000  pounds  of  steam  at  100  degrees  superheat, 
allowance  being  made  for  all  heat  returned  to  the  boilers. 


Fig.  83A.  —  Worthington  Three-stage  Turbine,  used  in  the  Montreal  Waterworks. 

Test  showed  that  the  pump  and  engine  reached  a  duty  of  113,302,278 
foot-pounds.  The  superheat  was  166  degrees. 

A  second  test  gave  a  duty  of  108,053,861  foot-pounds,  at  an  average 
of  107  degrees  superheat.  This,  with  superheat  at  100  degrees,  would  be 
equivalent  to  107,588,268  foot-pounds. 

A  third  test  showed  a  total  duty  of  113,557,732  foot-pounds  at  119.6 
degrees  superheat,  which  is  equivalent  to  110,151,000  foot-pounds  at 
100  degrees  superheat. 

The  reason  for  the  variation  in  the  results  is  that  the  first  test  took  place 
in  the  summer  when  the  feed  water  was  warm  and  second  test  took  place 
in  the  fall  when  the  water  was  very  cold,  and  the  colder  the  water  the 
greater  the  value  of  the  return  heat.  If  allowance  is  made  for  the  varia- 
tion in  superheat  and  feed  water,  the  three  tests  are  considered  to  check 
up  so  closely  as  to  preclude  errors. 


GENERAL   REMARKS 


87 


TESTS  OF  CENTRIFUGAL  PUMPING  UNIT  AT  MONTREAL. 


Second  Test. 

Third  Test. 

Hours  of  test                                        

24 

13 

Imperial  gallons  pumped             .      .  .        

12,042,000 

6,515,000 

Vvcrage  corrected  head,  feet."-.  

204.94 

204.98 

Work  per  hour,  foot-pounds  

1,027,772,000 

1,027,267,000 

Average  corrected  steam  pressure,  pounds  
Vvcrage  superheat  at  engine,  degrees 

144.4 
107  4 

142.7 
119  6 

\ir  pump  discharge  per  hour,  pounds 

9,673 

9,665 

Duty  per  1000  pounds  steam  condensed,  foot-pounds 
Duty  per  1000  pounds  steam  supplied,  corrected 
for  100  degrees  superheat,  and  for  heat  returned 
to  boiler   foot-pounds 

103,628,310 
107,588,268 

106,287,000 
110  151  000 

The  general  appearance  of  the  unit  is  shown  in  Fig.  83A.  The 
surface  condenser  is  located  in  the  suction  pipe  and  its  air  pump  is 
driven  by  a  belt  from  a  pulley  on  the  engine  shaft  near  the  coupling 
of  the  pump  shaft.  The  total  floor  space  occupied  by  the  unit  is  about 
23  by  40  feet. 

In  the  second  test,  it  was  found  impossible  to  keep  the  superheat  at  a 
uniform  temperature  on  account  of  the  existing  conditions  in  the  station. 
It  varied  considerably,  falling  below  100  degrees  a  number  of  times. 

In  the  third  test,  the  air  pump  discharge  was  weighed  on  calibrated 
scales,  and  the  weighing  was  checked  at  frequent  intervals  throughout 
the  trial.  It  was  not  possible  to  test  the  tightness  of  the  condenser,  but 
as  any  leakage  there  would  go  against  the  engine,  and  none  was  apparent, 
it  was  considered  advisable  not  to  consider  this  chance  of  error.  The 
steam  temperature  was  measured  by  a  Hicks  certified  thermometer  prob- 
ably correct  within  J  degree  Fahrenheit.  The  pump  suction  was  meas- 
ured by  a  mercury  column.  The  water  pumped  from  the  heater  to  the 
boiler  was  measured  by  a  meter  which  was  tested.  Feed  water  tempera- 
tures were  taken  with  tested  thermometers. 

Owing  to  a  burst  main  it  was  necessary  to  shut  down  the  pump  after 
running  13f  hours.  It  was  impossible  to  maintain  the  superheat  at  the 
specified  amount,  and  the  steam  consumption  actually  measured  was 
corrected  in  the  following  manner.  The  steam  consumption  per  pump 
horse-power-hour  and  the  average  superheat  for  each  hour  were  plotted 
on  a  diagram  together  with  the  guaranteed  steam  consumption  of  the 
engine.  From  this  it  appears  that  at  100  degrees  superheat  during  the 
test  the  steam  consumption  would  be  19.50  pounds  per  pump  horse-power- 
hour  as  against  18.82  pounds  at  120  degrees  superheat,  an  increase  of 
3.6  per  cent.  From  the  curve  based  on  the  guarantee  the  increase  per 
horse-power-hour  was  found  to  be  from  12.85  to  13.25  pounds,  an  in- 


88  CENTRIFUGAL  PUMPING  MACHINERY 

crease  of  3.1  per  cent.  It  was,  therefore,  considered  that  it  was  fair  to 
make  an  allowance  of  3  per  cent  for  the  increase  in  the  steam  consumption 
when  the  superheat  averaged  100  degrees  instead  of  119.6  degrees.  The 
actual  steam  consumption  as  measured  was,  therefore,  multiplied  by, 
1.03  to  allow  for  this. 


CHAPTER  XV. 
IRRIGATION,   DRAINAGE,  AND  SEWAGE. 

THE  work  here  is  particularly  well  suited  to  centrifugal  pumps,  as  the 
heads  are  low  and  the  capacities  large.  Foreign  substances  such  as  sand, 
silt,  and  sewage  matter  must  be  handled,  and  as  the  centrifugal  pump  is 


Fig.  84.     66-inch  Irrigating  Pump. 

entirely  free  from  valves  and  has  large,  easy  passages,  foreign  substances 
pass  through  without  choking  or  obstructing  the  pump.  Water  impreg- 
nated with  acids,  as  is  the  case  in  sewage,  does  not  tend  to  destroy  them, 


90 


CENTRIFUGAL  PUMPING  MACHINERY 


as  in  the  reciprocating  type,  which  offer  more  resistance  and  form  adverse 
currents  by  their  partitions,  valve  gratings,  and  valves. 

The  interior  of  a  centrifugal  pump  can  be  readily  coated  with  a  non- 
corrosive  enamel,  covered  with  hot  asphaltum  and  tar,  which  will  form  a 
solid  enamel  and  prolong  the  life  of  ordinary  cast  iron  to  that  of  bronze. 

In  drainage  work  the  prevailing  lifts  vary  from  3  to  50  feet.  It  is  well 
known  that  along  the  entire  Mississippi  River  a  great  deal  of  land  has  been 
reclaimed,  particularly  near  the  bayous  of  Louisiana  and  Texas,  where  the 
rice  crops  now  being  raised  do  not  depend  upon  the  rainfall.  Irrigation 


Constant  Capacity 
100,000  Gallons  per  Minute 


1  2  3  4  5  6  1  8  9  10          11          12          13 


Fig.  85.     Performance  Curve  of  66-inch  Irrigating  Pump. 

through  a  series  of  canals  now  employed  has  revolutionized  the  rice  in- 
dustry. The  amount  of  water  required  to  flood  these  fields  corresponds 
to  an  average  depth  of  2  feet  or  more  for  a  season. 

Egypt  has  again  come  to  the  front  as  a  producing  country,  due  to  the 
introduction  of  modern  centrifugal  pumps,  and  is  now  covered  with  a 
series  of  canals  supplied  by  large  irrigating  plants  installed  during  the  last 
five  years,  many  of  them  of  large  size,  ranging  from  18  to  60  inches.  The 
total  amount  of  water  these  plants  deliver  is  enormous. 

Irrigation  work  is  simply  the  reverse  of  drainage;  in  the  former  the 
water  is  lifted  and  distributed,  and  in  the  latter  it  is  drained  and  discharged. 

Fig.  84  illustrates  a  66-inch  irrigating  pump  for  100,000  gallons  per 
minute,  against  a  head  of  8J  feet  at  100  revolutions,  and  125,000  gallons 


IRRIGATION,  DRAINAGE,  AND  SEWAGE 


91 


against  a  head  of  about  20  feet  at  increased  maximum  speed  of  150 
revolutions;  the  former  for  75  per  cent  efficiency  and*the  latter  85  per  cent. 

Fig.  85  shows  the  performance  curve  of  this  pump  at  100  revolutions. 

Fig.  86  shows  the  installation  of  the  Jennings  Canal  Company,  Jennings, 
La.,  on  the  Bayou  Nez  Pique,  a  representative  type.  A  description  of 
this  plant  follows: 


Suet 


"~t_J 


Fig.  86.     Installation  of  a  Worthington  Centrifugal  Pump,  Plant  of  Jennings  Canal 

Company,  Jennings,  La. 

This  is  a  Worthington  centrifugal  pump  of  the  horizontal-shaft  volute 
single-suction  type;  the  suction  pipe  is  30  inches  in  diameter.  The  flume 
into  which  the  pump  discharges  is  built  into  the  pumping  plant  and  directly 
over  the  pump,  so  that  it  discharges  the  water  upward  through  an  expand- 
ing nozzle  30  inches  in  diameter;  the  top  of  the  nozzle  is  rectangular  in 
cross  section  and  more  than  twice  the  area  of  the  lower  end,  so  that  the 
velocity  of  water  at  discharge  is  greatly  reduced. 

The  normal  capacity  of  the  pump  is  22,000  gallons  per  minute,  but  can 
be  increased  to  a  much  higher  figure  by  an  increase  of  speed.  The  pump  is 
direct-connected  to  a  Hamilton  tandem  compound  condensing  four-valve 
engine,  size  12  inches  and  24-by-24-inch  stroke,  provided  with  a  shaft  gov- 
ernor. The  condenser  is  a  Deane  independent  jet. 


92 


CENTRIFUGAL   PUMPING  MACHINERY 


t,  14000 


A  Babcock  &  Wilcox  water-tube  boiler  is  installed  which  has  1550  square 
feet  of  heating  surface.  The  steam  pipe  is  short  and  5  inches  in  diameter. 
The  smokestack  is  95  feet  high  and  32  inches  in  diameter.  Feed  water  is 
supplied  by  a  Worthington  duplex  piston  pump  5J  by  3J  by  5  inches. 

A  Webster  175-horse-power  heater  raises  the  temperature  of  the  feed 
water  from  80°  F.  to  nearly  170°.  The  fuel  used  is  crude  oil,  costing  95 
cents  per  barrel  of  42  gallons,  or  320  pounds,  at  the  plant.  The  heat 
value  of  the  oil  is  18,922  B.t.u.  per  pound.  It  is  pumped  to  the  burner  by 
means  of  a  3-by-2-by-3-inch  duplex  pump.  The  burner  is  a  Peabody  No.  1. 
Some  preliminary  observations  were  made  to  find  the  economical  speed, 
to  adjust  the  engine  valves,  and  to  get  the  whole  plant  running  under  favor- 
able conditions.  The  main  test  lasted  for  four  hours,  during  which  time 
all  the  observations  were  of  great  uniformity.  As  the  boiler  was  of  the 

cross-drum  type  with 
small  water -storing 
capacity  the  length 
of  time  was  sufficient 
for  a  test,  which  was 
satisfactory  in  every 
way. 

The  quantity  of 
fuel  oil  measured  at 
each  reading  was  160 
pounds,  and  the  time 
to  consume  this 
amount  was  never 

O.OU  -X.UU  U.*»U  U.OU  I.OV  _  ,  jT\/-\  • 

less  than  28  minutes 

nor  more  than  29J  minutes.  Steam  pressure  was  controlled  mainly  by 
means  of  the  damper  after  the  proper  setting  of  the  valves  of  the  burner 
had  been  determined.  The  feed  water  was  carefully  weighed  in  barrels. 
At  half -hour  intervals  indicator  cards  were  taken,  together  with  the  general 
observations  given  in  the  log.  Steam  pressures  were  obtained  by  means 
of  calibrated  gauges.  The  amount  of  water  pumped  was  measured  in  the 
flume  by  means  of  a  Price  current  meter. 

A  calibrated  vacuum  gauge  was  attached  to  the  suction  pipe,  from  which 
the  suction  head  was  read.  Although  this  pipe  was  enlarged  at  the  lower 
end,  there  was  some  loss  of  head  due  to  entrance  loss  and  friction.  Above 
the  pump  an  opening  was  made  in  the  discharge  nozzle  near  the  flange, 
where  the  diameter  was  30  inches,  and  the  discharge  pressure  read  from  a 
glass  column  attached  by  means  of  rubber  tubing.  The  head  on  the  pump 
and  the  difference  of  levels  are  given  in  the  log. 

This  plant,  when  the  size  of  the  unit  is  considered,  made  a  splendid  show- 
ing. The  amount  of  fuel  oil  per  useful  water  horse  power  was  2.16  pounds. 


Tine 


IRRIGATION,  DRAINAGE,  AND   SEWAGE  93 

SUMMARY  OF  RESULTS. 
Test  of  Boiler,  Engine,  and  Pump,  Jennings  Canal  Company,  August  10,  1907. 

Lverage  steam  pressure  gauge 145. 4 

Lverage  temperature  of  water  from  heater,  degrees  Fahrenheit 169.  5 

Average  temperature  of  water  pumped 86. 

Lverage  temperature  of  flue 404. 

ition  of  boiler  test,  hours 4. 

Total  pounds  of  feed  water 18,594. 

Factor  of  evaporation 1.094 

Total  pounds  of  fuel  oil 1344. 

Amount  of  water  in  fuel  oil,  per  cent 1. 

Boiler  horse  power  (basis  34.5  evap.  from  and  at  212  degrees) 144. 

Ratio  of  water  evaporated  to  fuel  oil,  actual 13. 52 

Ratio  of  water  evaporated  to  fuel  oil  from  and  at  212  degrees 14. 8 

Efficiency  of  boiler  (18,922  B.t.u.  per  pound  of  fuel  oil),  per  cent 75. 5 

Average  rate  of  fuel  oil  per  24  hours,  barrels  of  320  pounds  or  42  gallons.  . .  25. 2 

Duration  of  engine  and  pump  tests,  hours 4. 

Average  vacuum  at  condenser,  inches  of  mercury 25. 

Average  revolutions  per  minute 187.  7 

Quality  of  steam  at  engine,  per  cent 97.  8 

Pounds  of  steam  used  by  engine,  vacuum  and  feed  pumps  and  burner  per 

I.H.P.  hour 20. 5 

Average  quantity  of  water  pumped,  gallons  per  minute 25,760. 

Average  head,  suction  level  to  discharge  level,  feet 24. 04 

Average  head,  including  friction  losses,  feet • 25.  75 

Average  I.H.P 221. 4 

Average  useful  water  horse  power 155. 7 

Efficiency  of  engine  and  pump,  per  cent 75. 4 

Efficiency  of  engine  pump  and  pipe  (including  friction),  per  cent 70. 3 

Price  of  fuel  oil,  per  barrel  of  42  gallons  or  320  pounds,  dollars .95 

Barrels  of  fuel  oil  per  hour 1 . 05 

Pounds  of  fuel  oil  per  minute 5.  60 

Heat  value  of  fuel  oil,  by  calorimeter,  B.t.u 18,922. 

Heat  value  of  fuel  oil  per  minute,  B.t.u 106,000. 

Heat  equivalent  to  I.H.P.  per  minute,  B.t.u 9370. 

Heat  equivalent  to  useful  water  horse  power  per  minute 6590. 

Ratio,  heat  equivalent  of  I.H.P.  to  heat  value  of  oil .  0883 

Ratio,  heat  equivalent  of  useful  horse  power  to  heat  value  of  oil .  0622 

Duty,  millions  of  foot-pounds  of  useful  work  per  million  B.t.u.  in  fuel.  .  .  48. 3 

Duty,  millions  of  foot-pounds  of  useful  work  per  1,000  pounds  of  steam. .  67. 7 

Cost  of  fuel  per  hour,  cents 100. 

Cost  of  fuel  per  acre  foot,  cents 21. 1 

Cost  of  fuel  per  foot,  acre  foot,  cents - .88 

Cost  of  fuel  to  raise  1,000,000  gallons  1  foot,  cents 2. 68 


94 


CENTRIFUGAL  PUMPING  MACHINERY 


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IRRIGATION,  DRAINAGE,  AND   SEWAGE 


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96  CENTRIFUGAL  PUMPING  MACHINERY 

Fig.  87  shows  the  Dry  Irrigating  pumps,  which  were  installed  by  the 
Madras  Government.     They  comprise   eight  39-inch  volute  pumps  de-  j 
signed  to  run  at  180  revolutions  per  minute,  direct-connected  to  vertical 
Diesel  engines.     Each  pump  will  pump  26,000  gallons  per  minute  against 
a  head  of  12  feet. 

Fig.  88  shows  the  type  of  irrigating  pumps  in  use  in  Colorado  and  places 
where  water  has  to  be  elevated  100  feet  or  more.     There  are  a  number  of 
these  installations  in  operation  for  the  Redlands  Irrigating  Company,  the 
Orchard  Construction  Company,  and  several  other  companies  which  are  i 
rendering  productive  land  which  was  formerly  considered  of  little  value. 

The  United  States  Government  has  introduced  the  policy  of  reclaiming 
land  in  the  arid  and  semiarid  regions  that  are  located  between  the  Missouri 
River  and  Rocky  Mountains,  and  about  8000  canals  and  3500  miles  of 
ditches  had  been  built  up  to  1909.  The  most  important  work  has  been 


Fig.  87.     Dry  Irrigating  Pump  at  Madras. 

done  in  the  North  Platte  district,  which  includes  a  part  of  Wyoming  and 
Nebraska;  the  Shoshone  district  in  Wyoming,  with  an  elevation  of  4000 
to  5000  feet;  the  Salt  River  district  in  Arizona;  the  Huntley  district  of 
Montana,  with  an  elevation  of  3000  feet;  the  Wittestone  district  of  North 
Dakota,  where  water  is  pumped  from  the  Missouri  River,  and  where  the 
banks  of  the  Missouri  are  so  exposed  to  changes  that  floating  pumping 
stations  were  made  necessary;  the  Buford  district,  North  Dakota;  the 
Garden  City  district  of  Kansas,  which  uses  an  underground  supply  for 
irrigation.  In  all  these  stations,  the  pumping  equipments  are  either  of 
the  motor-driven  centrifugal  or  of  the  gas-engine-driven  centrifugal  type. 
When  a  pumping  station  is  required  for  occasional  use  only,  the  equip- 
ment should  be  simple,  and  economy  is  of  secondary  importance.  For 
draining  or  irrigating  large  areas  of  land,  where  pumping  is  to  be  carried  on 
uninterruptedly  for  months  during  the  wet  or  dry  season,  there  should  be 
an  economical  installation,  with  carefully  designed  collecting  canals,  piping 
and  outlets,  so  that  water  is  not  pumped  even  a  few  inches  higher  than 
necessary. 


IRRIGATION,  DRAINAGE,  AND   SEWAGE 


97 


To  secure  an  economical  design,  the  pipes  should  be  arranged  to  form  a 
siphon,  with  the  pump  on  the  top  of  the  siphon.  The  ends  of  the  suction 
and  delivery  pipe  form  the  ends  of  the  siphon,  which  are  determined  by  the 
lowest  inner  and  outer  levels  reached  under  any  working  conditions.  Suc- 
tion pipes  should  be  short,  and  plenty  of  room  should  be  allowed  around 
them  so  as  to  prevent  currents  or  vortexes.  Water  should  enter  the  suc- 
tion pipe  quietly  and  equally  around  the  entire  rim.  The  minimum  diam- 


Fig.  88.     Type  of  Pump  used  in  Colorado  and  Places  where  Water  is  to  be  elevated 

100  feet  or  more. 


eter  of  the  suction  well  should  be  four  times  that  of  the  pipe,  and  when 
properly  designed  the  water  may  be  pumped  to  within  a  foot  of  the  rim  of 
the  pipe.  The  bottom  or  rim  of  the  suction  pipe  should  be  of  the  bell 
form,  known  as  vena  contracta,  which  gives  good  results  for  low  lifts. 

It  is  also  desirable  to  have  the  suction  pipe  gradually  enlarge  from  the 
pump,  so  that  the  high  velocities  will  be  reached  gradually.  No  screen 
should  be  fitted  on  the  suction  pipe  itself,  but  one  should  be  placed  across 
the  canal  or  well  to  prevent  foreign  matter  from  entering  the  pump. 


98 


CENTRIFUGAL  PUMPING  MACHINERY 


The  delivery  pipes  should  be  of  gradually  expanding  section,  so  that  the 
velocities  at  the  point  of  discharge  are  reduced  to  from  4  to  2  feet  per 
second,  and  should  discharge  into  a  still  pool.  This  change  of  diameter  in 
the  pipe  should  take  place  in  not  less  than  8  to  12  diameter  lengths,  other- 
wise the  velocities  will  be  changed  too  abruptly. 

The  remarks  about  irrigating  and  drainage  pumps  and  the  description 
of  types  apply  in  a  great  measure  to  pumps  for  sewage,  of  which  there  are 
many  notable  installations  in  such  cities  as  Boston,  Chicago,  New  Orleans, 
and  Milwaukee,  which  give  excellent  results  and  good  economy. 


Fig.  89.     Arrangement  of  Horizontal  Sewage  Pump  installed  by  the  Belfast  Cor- 
poration, England,  at  the  Greencastle  Pumping  Station. 


Fig.  89  shows  the  arrangement  of  horizontal  sewage  pumps  installed  by 
the  Belfast  Corporation  at  the  Greencastle  Pumping  Station,  Belfast,  Eng- 
land. In  this  case  most  of  the  pumping  head  is  on  the  suction  side,  the 
pumps  being  automatically  controlled  by  the  suction  level.  Where  vertical 
pumps  are  necessary,  the  type  shown  in  Fig.  90  is  used,  which  allows  natural 
priming.  Care  should  be  taken  in  the  design  of  pumps  for  unscreened 
sewage,  as  special  impellers  and  large  casings  are  necessary,  and  storm 
relief  pumps  must  be  so  designed  that  they  can  handle  considerable  sand 
and  grit. 


IRRIGATION,  DRAINAGE,  AND  SEWAGE 


99 


Fig.  90.    Arrangement  of  Vertical  Pumps  at  the  Greencastle  Pumping  Station, 

England. 


CHAPTER  XVI. 
HYDRAULIC   MINING   AND   DREDGING. 

THE  removal  of  earth  in  large  quantities  can  be  accomplished  more 
quickly  and  cheaply  by  means  of  centrifugal  pumps  than  with  steam 
shovels.  This  is  being  done  on  the  Pacific  Coast  in  various  places,  both 
for  mining  and  grading,  and  also  at  the  phosphate  mines  in  Florida. 

To  remove  the  loosened  material  special  dredging  pumps  are  usedj 
which  are  fitted  with  manganese  impellers.  The  casings  may  be  lined  or- 
unlined,  depending  upon  the  size  of  the  pump  and  the  nature  of  the  work 
to  be  done.  Such  pumps  are  capable  of  handling  a  large  amount  of  sus- 
pended solid  matter,  even  up  to  25  per  cent  of  the  water,  and  are  designed 
so  as  to  take  care  of  stones  4  to  12  inches  in  diameter. 


Fig.  91.     10-inch  Ten-stage  Hydraulic  Mining  Pump. 

Fig.  91  shows  one  of  the  four  10-inch  ten-stage  pumps  installed  in  Seattle 
for  grading  Jackson  Street,  used  in  removing  two  million  cubic  yards  of 
earth. 

The  Panama  Canal  Commission  have  installed  centrifugal  pumping 
engines  to  remove  7,500,000  cubic  yards  of  material,  consisting  of  soft  silt, 
earth,  clay,  and  hard  rock.  Various  attempts  were  made  with  dredges,  but 
the  final  conclusion  was  to  handle  it  by  hydraulic  methods.  The  material 
to  be  loosened  and  moved  by  the  dredging  pumps  is  dark  loam,  containing 
15  per  cent  sand  and  gravel,  and  weighs  75  pounds  per  cubic  foot. 

Centrifugal  pumps  are  frequently  used  for  dredging  and  pumping  mud 
from  rivers  and  harbors.  The  cost  of  operating  centrifugal  dredging 
pumps  compares  favorably  with  ladder  or  bucket  dredges,  and  in  some 
cases  is  more  economical.  A  modification  of  this  pumping  of  suspended 

100 


HYDRAULIC   MINING  AND   DREDGING 


101 


material  is  found  in  wrecking  work,  where  hulls  are  cleared  of  wheat  and 
[other  substances  which  have  no  cohesion. 

Centrifugal  dredgers  have  been  employed  extensively  by  the  United 
States  Government  for  deepening  rivers  and  harbors,  and  filling  in  and 
reclaiming  land  by  pumping  the  dredged  material  through  piping.  Such 
work  requires  velocities  in  the  pipes  of  about  10  feet  for  about  3  to  5  per 
cent  of  solid  material,  which*  consists  of  rock,  clay,  sand,  shells,  and  mud. 
Discharge  piping  is  made  of  either  wood  or  metal,  wood  giving  the  better 
results  in  regard  to  wear  and  tear. 


Fig.  92.     Horizontal  20-inch  Wilredging  Pump.     Horizontal  Steel  Lincel  Type. 

Fig.  92  shows  the  large  dredging  pumps  installed  at  Panama  by  the  Canal 
Commission.  There  are  four  of  the  horizontal-shaft  type,  each  with  a  sin- 
gle 20-inch  side  suction  and  a  20-inch  delivery,  and  are  exceedingly  heavy 
in  construction.  The  runners  or  impellers  are  of  manganese  steel  and  of 
the  inclosed  type.  The  entire  internal  surface  of  the  casing  is  covered 
with  soft  boiler  plate,  and  the  casings  are  of  extra  thickness  to  stand  the 
heavy  work.  The  suction  openings  are  protected  by  removable  cast-steel 
throats  or  rings.  This  construction  makes  one  of  the  best  arrangements 
for  easy  repair.  The  shafts  are  nickel  steel. 

Each  pump  is  designed  to  deliver  10,000  gallons  of  water  per  minute, 
against  a  height  of  60  feet,  exclusive  of  pipe  losses  and  suction  heads  of 
10  feet  through  1200  feet  of  pipe  line.  A  suitable  thrust  bearing  of  marine 
type  is  provided  to  take  care  of  the  unbalanced  pressure. 


102  CENTRIFUGAL  PUMPING  MACHINERY 

Each  is  capable  of  excavating  and  disposing  of  300  cubic  yards  of  solid 
matter  per  hour,  using  10,000  gallons  of  water  per  minute,  or  about  10 
cubic  feet  of  water  per  cubic  foot  of  material,  which  is  loam  with  15  per 
cent  sand.  With  3000  horse  power  available,  the  efficiency  will  reach 
approximately  60  per  cent.  On  the  basis  of  10  cubic  feet  of  water  per 
cubic  foot  of  material,  and  estimating  the  85  per  cent  of  dirt  at  80  pounds 
per  cubic  foot,  and  the  15  per  cent  of  sand  at  110  pounds  per  cubic  foot, 
we  obtain  68  pounds  and  16J  pounds  respectively,  or  a  total  of  84  J  pounds 
per  cubic  foot.  Assuming  a  mixture  of  40  per  cent  clay  at  120  pounds  per 
cubic  foot,  giving  40  pounds,  15  per  cent  sand  at  16J  pounds,  and  45  per 
cent  loam  at  45  pounds,  we  have  an  average  weight  of  101|  pounds  for  the 
material.  Taking  the  weight  at  110  pounds  per  cubic  foot,  and  the  volume 
at  300  cubic  yards  per  hour,  the  material  equals  891,000  pounds  per  hour, 
or  14,850  pounds  per  minute,  plus  10,000  gallons  or  84,042  pounds  of 
water,  making  a  total  of  98,892  pounds  of  material  to  be  pumped  per 
minute  by  three  pumps;  the  fourth  being  a  " booster  "  pump.  The  total 
of  98,892  pounds  divided  by  10,000  gallons  will  give  9T9o  pounds  of  weight 
per  gallon,  or  74J  pounds  per  cubic  foot  of  mixture.  This  gives  10  per  cent 
of  solid  matter  by  volume  and  15  per  cent  by  weight. 

Some  of  the  mixtures  to  be  carried  weigh  90  pounds  per  cubic  foot, 
which  equals  a  weight  of  12  pounds  per  gallon,  or  a  total  of  120,000  pounds 
per  10,000  gallons.  Deducting  the  84,042  pounds  of  water  leaves  25,958 
pounds  of  solid  material,  or  about  21  per  cent  by  weight. 

These  figures  give  an  idea  of  the  hard  service  under  which  these  pumps 
operate.  Each  is  driven  by  a  655-horse-power  motor  of  the  two-bearing 
type,  running  at  480  revolutions  and  fitted  with  autostarters  and  autotrans- 
formers.  There  are  also  four  motors  operating  the  vacuum  pumps  of  the 
priming  equipment.  These  vacuum  pumps  are  of  the  single  horizontal 
type  operated  by  a  5-horse-power  motor.  There  is  also  a  complete  set  of 
transformers  and  marine  switchboard  for  starting  automatically,  fitted 
with  the  necessary  wattmeters,  oil  switches,  and  overload  trip  arrange- 
ment. There  are  all  the  accessories  for  starting  the  motor  on  the  compen- 
sator and  cutting  it  out  when  the  switches  are  in  running  position.  The 
current  for  operating  the  oil  switches  is  25  cycles,  stepped  down  from  2080 
volts  to  110  volts.  The  motor  connections  are  brought  to  a  box  near  the 
base  at  the  side  of  the  motor.  All  of  the  leads  are  insulated  and  bushed 
to  insure  safety  in  case  of  accidental  flooding. 

The  six  hydraulic  giants  or  monitors  are  of  the  latest  type  used  in  mining 
operations  in  California  and  other  parts  of  the  West.  Each  weighs  1500 
pounds  and  consists  of  a  base  for  attachment  to  a  16-inch  gate  valve  at 
the  terminus  of  the  pipe  line,  a  horizontal  and  vertical  joint,  and  a  long, 
conical  reducing  piece.  The  frictional  resistance  is  decreased  by  a  ball 
bearing,  and  a  weighted  lever  is  attached  to  control  the  direction  of  the 


HYDRAULIC   MINING  AND  DREDGING  103 

jet.  A  deflecting  nozzle  is  fitted  to  the  discharge  end  of  the  giant,  which 
permits  deflections  through  a  small  angle  without  changing  the  position  of 
the  main  body.  The  tapering  piece  of  the  giant  is  fitted  on  the  inner  side 
with  two  sets  of  guide  vanes  which  prevent  a  scattering  or  rotary  motion 
of  the  water  after  it  has  issued  from  the  nozzle.  The  nozzles  used  vary 
from  4  to  6  inches  in  diameter,  according  to  the  character  of  the  material  in 
which  they  are  working,  and  at  full  head  the  water  coming  through  them 
exerts  a  pressure  of  130  pounds  to  the  square  inch,  the  equivalent  of  a  ton 
and  one-half  of  pressure  against  a  bank  100  feet  away  within  range  of  the 
deflectors.  As  it  is  expected  that  the  positions  of  the  monitors  will  be 
shifted  frequently,  their  bases  are  of  temporary  construction.  The  areas 
excavated  and  filled  by  the  giants  are  8  feet  above  mean  tide  and  the 
average  depth  to  be  excavated  is  45  feet.  This  is  accomplished  by  wash- 
ing down  the  material  in  sluices  which  carry  the  water  containing  earth  in 
suspension  to  the  sump  where  the  barge  or  dredging  pumps  are  at  work. 
When  it  becomes  necessary  to  move  a  barge,  the  giants  cut  a  new  sump 
with  a  channel  leading  into  it  through  which  one  or  more  units  of  the  fleet 
is  floated.  The  banks  are  excavated  as  nearly  perpendicularly  as  possible, 
in  order  that  benches  may  be  cut  in  them  and  the  banks  undermined  so  as 
to  cause  the  material  to  fall  by  gravity. 


CHAPTER  XVII. 
MINING   WORK. 

THE  turbine  pump  has  been  extensively  used  for  removing  water  from 
mines,  for  station  or  drainage  service,  and  for  sinking. 

Fig.  93  shows  an  8-inch  eight-stage  pump  for  a  lift  of  1250  feet.  Sev- 
eral of  these  pumps  are  in  successful  operation  handling  large  quantities  of 


Fig.  93.     8-inch  Eight-stage  Pump.     Deep  Mining  Pumps. 

water,  each  impeller  operating  against  as  high  as  157  feet,  the  efficiency 
being  75  per  cent. 

In  a  turbine  type  of  pump  it  is  practicable  to  have  single  lifts  as  high  as 
1600  to  1800  feet.  In  reciprocating  pumps  the  lifts  rarely  exceed  1000  to 
1200  feet.  Several  large  pumps  handling  5000  to  6000  gallons  per  minute, 
against  500  feet  lift,  have  been  in  successful  operation.  For  acid  water  the 
working  parts  are  made  of  special  acid-resisting  bronze,  containing  copper, 
tin,  and  lead. 

The  use  of  electric  sinking  pumps  is  becoming  general,  particularly  for 
emptying  mines  that  are  flooded.  In  sinking  work  a  simple  suction  pump 
is  provided,  so  that  the  suction  pipe  cannot  be  uncovered.  This  allows 
the  pump  to  operate  continuously  with  an  ample  supply  of  water. 

Figs.  94  and  95  illustrate  sinking  pumps  for  deep  working.  The  use  of 
this  type  eliminates  all  steam  and  exhaust  pipes.  It  is  to  be  noted  that  a 
sinking  pump  should  be  proportional  for  the  final  head  against  which  it 
has  to  pump,  and  special  precautions  must,  therefore,  be  exercised  in  its 
design.  The  pumps  are  either  self-supporting  or  hung  in  frames.  This 
freedom  from  shocks,  particularly  in  freely  suspended  pumps,  allows  high 
velocities  of  water,  thereby  reducing  the  size  of  pipe  column. 

104 


MINING  WORK 


105 


By  means  of  specially  designed  gate  valves  in  the  discharge,  the  quantity 
of  water  delivered  can  be  regulated  to  the  flow  into  the  pump,  and  by  a 
special  design  of  suction  chambers  the  accumulation  of  air  can  be  prevented. 


IMC 


Figs.  94  and  95.     Sinking  Pumps  for  Deep  Working. 

The  advantage  of  a  centrifugal  sinking  pump  lies  in  its  compactness  and 
it  weight,  with  the  consequent  facility  in  handling  and  transporting  in 
shafts. 


106  CENTRIFUGAL  PUMPING  MACHINERY 

In  introducing  the  turbine  pump  for  mining  purposes,  considerable  prej- 
udice will  have  to  be  overcome,  on  account  of  the  unsatisfactory  results 
which  have  been  due  to  wrong  selection  of  pumps,  and  to  the  fact  that 
manufacturers  have  not  studied  the  requirements  properly.  The  voltage 
may  vary  from  outside  causes,  and  a  motor-driven  turbine  pump  should  be 
so  installed  that  it  will  be  reliable  under  all  conditions. 

If  the  speed  of  the  motor  and  pump  under  the  fluctuations  of  the  line 
voltage  and  variation  of  capacity,  and  also  the  power  of  the  motor  against  a 
constant  head,  be  carefully  considered,  and  the  pump  designed  accordingly, 
there  is  no  reason  why  it  should  not  be  used  satisfactorily.  The  efficiencies 
obtainable  are  dependent  upon  the  capacities  for  the  heads  under  which 
the  pump  is  to  be  installed,  and  attention  should  be  paid  to  that  point  if 
good  results  are  to  be  expected.  With  the  proper  relation  between  capacity 
and  head,  efficiencies  equal  to  reciprocating  power  pumps  may  be  expected, 
and  with  provisions  for  taking  care  of  the  fluctuations  in  the  line,  centrifugal 
pumps  should  meet  with  no  difficulties. 


CHAPTER  XVIII. 
POWER-STATION  WORK. 

Qjfi^    BOILER  FEEDING. 

IT  is  only  lately  that  this  promising  field  for  the  multistage  turbine  pump 
has  been  invaded.  The  turbine  pump  is  well  adapted  to  medium  and  large 
power-station  work,  where  it  will  furnish  an  uninterrupted  and(  continuous 
flow! of  water,  free  from  shocks  and  water  hammers,  obviates  the  otherwise 
necessary  (air  chambers,  relief  valves,)  and  has  a  distinct  advantage  over  the 
reciprocating  pump  in  point  of 'lubrication  and  attendance.'  The  (life  of 
both1  piping  and  pumps  Vill  be  prolonged, -and  the  plant  will  be  increased 
in  efficiency."  It  is  advisable  in  large  power  stations,  however,  to  have  a 
stand-by  pump  of  the  reciprocating  type  to  be  used  in  connection  with  the 
turbine  pumps.  There  is  no  possibility  of  building  up  undue  pressure  by 
closing  the  discharge  or  feed  valves.  No  trouble  will  come  from  steam- 
pressure  regulators  and  safety  valves  if  properly  installed,  but  judgment 
must  be  used  in  relation  to  the  drop  in  steam  pressure,  to  low  water,  and 
peak  loads;  and  a  reciprocating  pump  must  be  kept  for  emergency  and  to 
take  care  of  variations  in  capacity  and  irregularities. 

f  Centrifugal  pumps,)  if  motor-driven  under  constant  speed,  (should  be 
designed  for  a  considerable Jrange  of  delivery  that  increase  of  pressure  will 
be  taken  care  of,  with  corresponding  capacity.  In  a  steam-turbine-driven 
pump  this  can  be  accomplished  by  increased  speed.  A  motor-driven  pump 
can  also  be  arranged  for  variation  in  speed)if  desired.  There  is  no  fear  of 
accident  with  a  turbine  pump,  whether  motor-  or  steam-turbine-driven. 

Quite  a  number  of  these  pumps  are  installed  in  large  power  stations  like 
the  Commonwealth  station  at  Chicago,  the  New  York  Central  power  station 
at  Xew  York,  the  Interboro  power  station,  and  others,  and  abroad  in  the 
Edison  Milano  Univeraria  Valdarno,  in  Turin,  Milan,  Florence,  Birming- 
ham, Amsterdam,  and  other  places. 

Fig.  96  illustrates  a  steam-turbine  boiler-feed  pump  for  250  pounds 
pressure.  Fig.  97  shows  a  steam-turbine-driven  pump  for  350  pounds 
pressure.  Fig.  98  illustrates  a  motor-driven  boiler-feed  pump  for  250 
pounds  pressure. 

In  boiler-feed  pumps  it  is  particularly/desirable  to  have  perfectly  balanced 
conditions  and  free,  uninterrupted  flow  to  impellers,  with  velocities  as  low 

107 


108  CENTRIFUGAL  PUMPING  MACHINERY 


Fig.  96.     Steam-turbine  Boiler-feed  Pump. 


Fig.  97.     Steam-turbine-driven  Boiler-feed  Pump. 


Fig.  98.     Motor-driven  Boiler-feed  Pump. 


POWER-STATION   WORK 


109 


as  possible.  J  An  excellent  design  is  shown  in  Fig.  99,  where  this  is  accom- 
plished by  admitting  the  water  simultaneously  to  two  impellers  turned 
back  to  back,  which  discharge  it  into  a  central  double-entrance  impeller. 
Good  engineering  requires  the  use  of  the  simplest,  most  durable,  and  most 
economical  apparatus.  Since  this  type  of  pump  has  been  introduced  there 
is  a  tendency  to  do  away  with  cumbersome  and  costly  power  pumps  and 
steam  reciprocating  pumps,  as  the  turbine  pumps  now  installed  show 

!— 10*-- 1 


Fig.  99.     Plan  of  Boiler-feed  Pump  with  Opposed  Balanced  Impellers. 

remarkable  results  for  high  speed,  low  cost  of  maintenance,  reliability  and 
smoothness  of  operation,  economy  in  space,  low  cost  of  installation,  and 
easy  attendance. 

CIRCULATION  OF  WATER. 

In  order  to  obtain  circulating  water  for  condenser  plants,  it  is  usual  to 
employ  a  centrifugal  pump  on  account  of  its  high  efficiency  under  the  con- 
ditions which  call  for  large  volumes  of  water  under  low  heads.  Several 
designs  are  used,  either  engine-,  motor-,  or  steam-turbine-driven,  having  a 
single  or  double  suction,  depending  upon  the  conditions. 

Condenser  overloads  should  be  taken  into  account,  and  the  capacity  of 
the  pumps  should  be  great  enough  to  take  care  of  the  extra  work.  The 
connections  to  the  pump  from  the  condenser  should  be  as  short  as  possible, 
with  no  air  pockets.  Means  should  be  devised  for  priming  the  pump,  or  a 
small  connection  may  be  made  to  the  vacuum  pump. 

In  engine-driven  pumps  the  regular  type  of  volute  design  of  pump  cham- 
ber and  impeller  can  be  easily  applied,  as  the  speed  is  normal.  The  siphon 
arrangement  as  used  hi  irrigating  work  is  most  desirable.  By  specially 


110 


CENTRIFUGAL  PUMPING  MACHINERY 


designed  high-speed  impellers  this  type  can  be  used  with  motor  or  steam 
turbine  drives  within  certain  limits  (see  Fig.  100) .  But  to  meet  the  higher 
speeds  of  steam  turbines  and  to  handle  large  quantities  of  water  under 
small  heads,  usually  only  friction  heads,  a  new  birotor  or  trirotor  pump 
has  been  placed  on  the  market.  These  pumps,  described  in  Chapter  XXIX, 
can  be  made  in  sizes  and  speeds  to  run  with  the  average  steam  turbine. 

In  large  condenser  installations  where  all  the  auxiliaries  are  to  be  driven 
by  steam  turbines,  a  new  type  of  centrifugal  vacuum  pump  of  the  jet  style 


Fig.  100.     Steam-turbine-driven  Pump  with  High-speed  Impellers. 

has  been  developed.  Several  of  these  are  being  built,  arranged  on  the 
same  shaft  with  the  circulating  and  hot-well  pumps,  making  a  very  com- 
pact unit,  the  three  auxiliaries  being  driven  by  one  steam  turbine. 

HOT-WELL  PUMPS. 

Closely  allied  to  the  boiler-feed  pump  is  the  hot-well  pump,  designed  for 
both  steam-turbine  and  motor  drive.  It  is  rarely  made  more  than  two- 
stage,  and  has  an  additional  vapor  opening  on  the  suction  entrance.  These 
pumps  withdraw  the  water  of  condensation  from  the  condenser.  They 
may  be  either  of  the  vertical  or  horizontal  type,  as  most  suitable  to  the 
particular  installation. 

Fig.  101  shows  a  pump  inclosed  in  the  hot  well,  making  a  very  compact 
arrangement.  Fig.  102  shows  the  regular  arrangement. 

The  turbine  hot-well  pump  has  practically  replaced  the  reciprocating 
type  in  all  modern  power  stations,  and  has  the  advantage  of  being  auto- 
matic in  action,  taking  care  of  all  the  variations  in  the  loads  without  any 


POWER-STATION   WORK 


111 


attention,  and  is  always  in  condition  to  operate  without  priming.    These 
pumps  are  direct-connected  to  either  a  steam  turbine  or  motor.     It  is  usual 


Fig.  101.     Hot  Well  Pump  Inclosed  in  Special  Well. 

to  locate  them  three  or  four  feet  below  the  water  line  in  condenser  in  order 
to  obtain  a  good  head  on  the  pump.     A  type  of  the  vertical  pump  installed 


Fig.  102.    Two-stage  Hot  Well  Pump. 

aboard  ship  is  shown  in  Fig.  103,  illustrating  the  compactness  and  lightness 
required  in  such  installations. 


112 


CENTRIFUGAL  PUMPING  MACHINERY 


CENTRIFUGAL- JET  CONDENSERS. 

The  development  of  steam  turbines  requiring  high  vacuum  has  obliged 
the  designers  of  centrifugal  pumps  to  provide  apparatus  which  will  give  a 

vacuum  within  one-half  inch  of  the 
absolute  vacuum.  Barometric  con- 
densers have  been  used  more  or  less 
for  this  work,  but,  owing  to  the  ex- 
tremely long  piping  with  consequent 
air  leaks,  they  have  given  way  to  the 
centrifugal  jet. 

Fig.  104  illustrates  an  engine- 
driven  unit  which  can  also  be  driven 
by  a  steam  turbine  or  motor.  It 
contains  'some  new  features  in  the 
design  of  the  working  parts  of  the 
pump,  and  in  the  condenser,  and 
may  be  built  either  on  the  two-stage 
or  single-stage  principle  of  the  pump, 
depending  upon  the  conditions.  The 
principle  involved  is  simple  and 
offers  the  designer  full  range  for 
his  ability  for  getting  good  results 
with  few  working  parts.  The  water 
is  taken  from  the  bottom  of  the  con- 
denser and  led  to  the  center  of  the 
impeller,  when  the  condenser  is 
alongside  of  the  pump.  When  the 
condenser  is  on  top  of  the  pump,  the 
water  is  guided  to  each  side  of  the 
impeller.  The  removal  of  the  non- 
condensable  vapors  is  accomplished 
either  by  a  rotative  dry-vacuum 
or  by  a  centrifugal  vacuum  pump. 
From  the  illustration  it  can  be  seen 
that  compactness  and  space  have 
been  considered  of  first  impor- 
tance. 

Figs.  105  and  106  show  the  ar- 
rangement for  vertical  and  horizontal 
jet  condensers,  with  trimotor  and 
birotor  pumps,  driven  by  steam  tur- 


Fig.   103.     Steam-turbine-driven  Vertical 
Marine  Hot  Well  Pump. 


bine,  motor,  or  engine.     Figs.  107  and  108  show  a  centrifugal  air-jet  conden- 
ser in  which  cold  water  for  the  condensation  of  the  steam  is  forced  through  a 


Fig.  104.     Engine-driven  Pump  Unit. 

.Air  Suction 

— n 

-Injection 


Fig.  105.     Arrangement  for  Vertical  Jet  Condensers.  (113) 


114 


CENTRIFUGAL  PUMPING  MACHINERY 


Air  Suction 


Fig.  106.     Arrangement  for  Horizontal  Jet  Condenser. 


Fig.  107.     Centrifugal  Water  Jet  Condenser  with  Centrifugal  Air  Exhausting  Pump. 


POWER-STATION   WORK 


115 


jet  nozzle,  absorbing  all  the  vapor  contained.  This  method  of  extracting 
the  air  dispenses  with  the  usual  dry-vacuum  pump.  The  water  and  con- 
densed vapors  are  removed  by  a  circulating  pump  in  such  a  manner  that 


Air  Pump 
Discharge 


Fig.    108. 


Section  of  a  Centrifugal  Water  Jet  Condenser  with  Centrifugal  Air 
Exhausting  Pump. 


the  air  is  drawn  off  by  the  outside  impeller  vanes  of  the  circulating  pump. 
This  construction  allows  direct  pumping,  and  obviates  a  separate  air  or 
vacuum  pump,  and  so  produces  a  compact  unit  of  low  cost,  high  efficiency, 
and  minimum  expense  for  operation  and  maintenance. 


CHAPTER  XIX. 
DOCKS. 

ONE  of  the  most  extensive  and  important  uses  of  the  centrifugal  pump  is 
its  application  to  pumping  out  docks,  and  it  is  here  that  the  largest  indi- 
vidual units,  ranging  from  36-  to  72-mch  discharge  opening,  are  used. 

This  service  calls  for  the  handling  of  large  volumes  of  water  in  the  least 
possible  time  under  heads  varying  from  0  to  approximately  48  feet  for  the 
drainage  pumps,  and  0  to  40  feet  for  the  main  pumps.  Great  difficulties 
are  met,  as  alternating  current  is  largely  used,  necessitating  constant  speed 
and  constant  horse  power  over  the  entire  range  of  capacity  and  head.  A 
great  many  pontoon  and  floating  dry-docks  have  been  equipped  with  this 
type  of  pump,  and  have  shown  that  the  time  taken  to  pump  out  docks  may 
be  lowered  considerably,  and  that  the  electric  power  required  may  be  re- 
duced about  40  per  cent  below  that  which  was  used  by  the  old-style  pumps. 

It  is  to  be  noted  that  in  a  dry-dock  the  sectional  area  decreases  as  the 
lift  increases,  and  that  the  most  economical  heads  will  be  found  from  about 
one  foot  above  keel  block  to  the  bottom  of  dock. 

The  largest  volumes  are  pumped  at  the  smallest  heads,  and  vice  versa, 
therefore  the  average  efficiency  is  very  important,  and  in  considering  this 
average  efficiency  note  should  be  made  between  what  levels  it  is  to  be 
considered. 

The  capacity  is  usually  judged  on  the  basis  of  an  average  between  zero 
head  at  mean  high-water  elevation  and  an  elevation  1  foot  above  the  top 
of  the  keel  blocks,  which  is  usually  from  4  to  6  feet  above  the  bottom  of  the 
dock.  The  keel  block  having  the  greatest  elevation  is  the  one  usually 
selected  as  the  pumping  point,  correction  being  made  for  tidal  variations. 
In  this  way  the  results  are  reduced  to  a  condition  of  constant  mean  high- 
water  level  outside  the  caisson  during  the  operation  of  pumping. 

Dock  pumps  are  of  both  the  vertical  and  horizontal  type,  the  latter  being 
the  more  common.  The  efficiencies  in  dry-dock  main  pumps  should  be 
judged  on  the  average  operating  conditions  given,  and  should  be  the  ratio 
of  the  work  done  in  pumping  against  the  head  to  the  power  input  of  the 
motor  measured  at  the  motor  terminal.  The  head  for  the  main  pump  is 
usually  taken  from  zero  at  mean  high- water  level,  encountered  when  the 
level  of  the  water  in  the  dock  is  1  foot  above  the  keel  blocks. 

The  efficiency  of  a  well-designed  dock  equipment  will  reach  about  42  per 
cent  rated  on  the  above  basis,  and  about  45  per  cent  from  the  bottom  of 

116 


DOCKS 


117 


the  dock,  the  pumping  unit  being  charged  with  all  f Fictional  losses  in  piping, 
valves,  etc.,  with  a  motor  efficiency  of  92  per  cent  and  from  85  to  90  per  cent 
efficiency  for  the  pumps  themselves.  Since  all  such  pumps  are  required 
to  run  only  for  short  periods  and  at  long  intervals,  efficiency  is  of  small 
importance,  and  the  first  consideration  is  reliability.  Pumps  are  generally 
furnished  in  duplicate.  It  can  be  assumed  that  a  dock  is  emptied  2000 


Fig.  109.     Performance  Curves  of  54-inch  Dry-dock  Pump. 

times  in  ten  years.  With  this  as  a  basis,  each  per  cent  of  increased  efficiency 
would  show  the  value  of  the  pumps  of  high  efficiency,  as  demonstrated 
towards  the  end  of  this  chapter. 

The  two  types  of  docks  are  the  graving  or  dry  dock  and  the  floating  or 
pontoon  dock.  The  former  has  a  permanent  pumping  station  located  at  a 
convenient  place  near  the  dock.  The  station  is  usually  in  a  depressed 
chamber  in  the  ground,  so  that  no  projections  are  encountered  in  handling 
the  hawsers  of  the  ships.  The  arrangement  of  piping  is  usually  made  as 
short  as  possible  to  avoid  losses. 

The  floating  pontoon  dock  is  of  a  somewhat  different  character,  and  may 
be  either  in  sections  or  in  one  unit,  arranged  so  that,  when  a  vessel  is  to  be 
clocked,  the  pontoon  is  submerged  by  admission  of  water  into  its  compart- 
ments to  pass  it  under  the  vessel.  The  pumps  then  empty  the  dock  com- 
partments until  the  vessel  and  dock  rise  sufficiently  to  expose  the  vessel 
so  that  work  can  be  done. 

The  centrifugal  pump  has  always  been  considered  ideal  for  such  pur- 
poses, and  the  low  cost  of  installation,  when  compared  with  that  of  recipro- 
cating pumps,  makes  its  use  universal. 


118 


CENTRIFUGAL  PUMPING  MACHINERY 


The  motive  power  may  be  either  a  steam  engine  or  electric  motor,  depend- 
ing upon  the  installation.     Engineers  usually  adopt  an  electrically  driven 


c  «-a 1  Pump  Based  on  Total  Head  (Bott 

2       "          "       "      "         "    (1  Foot  Above  Top  >f  Ke  1  Bio 
S   Pump  and  Motor  Based  on  Static  Head  (Bottom  of  Dock) 


ige  'Efflcieuc 


JU1 


Depth  of  Water  Dock 

Fig.  110.     Performance  Curves  of  45-inch  Dry  Dock  Pump. 

pump  on  account  of  the  convenience  with  which  current  can  be  had  for 
operating. 

The  time  taken  for  emptying  large  docks  varies  from  1£  to  2£  hours.    Fig. 


1.  Pump  based  on  Total  Head  (Bottom  (jf  Dock) 

?L_? I"    '1    "  I    " 


Average  Efficiencies: 


-J !- 


I     (1  Ift.above  Toty  of  Keel  Block) 


ip  ijMoto    based  on  Static  Head   |(Bottc|ni  of  Dock)  | 
"      "         "       "        "          "      .(I  Ft.above  JTop  of  Kee 


35  20  16 

-Depth  of  Water  in  Dock 


Fig.  111.     Performance  Curve  of  36-inch  Dry  Dock  Pump. 

109  shows  the  curves  of  a  54-inch  pump,  Fig.  110  of  a  45-inch  pump,  and 
Fig.  Ilia  36-inch  pump.     These  show  the  changes  in  pumping  conditions 


DOCKS 


119 


and  the  variation  in  capacities  and  heads  under  constant  speed  with  the 
almost  constant  horse  power  necessary  to  prevent  overload  on  the  motors. 
It  is  evident  from  an  examination  of  these  figures  that  the  problem  is  a 


/ 


/ 


\ 


Suction  Head          Impeller 


Casing 


Fig.  112.     Velocities  in  Single-entrance  Impellers  for  Dry-dock  Work. 

difficult  one,  as  the  design  of  the  impeller  must  take  into  account  the  varia- 
tion in  area  of  the  dock  at  its  various  levels,  and  the  outside  tide-water 
levels,  and  at  the  same  time  the  varying  power  required  from  no  head  to  full 


Suction  Head  Impeller  Casing 

Fig.  113.     Velocities  in  Double-entrance  Impellers  for  Dry-dock  Work. 

head  must  not  overload  the  motor.     The  curves  show  how  it  can  be 
accomplished. 

Vertical  pumps  having  one  suction  entrance  may  not  give  as  good  effi- 
ciencies as  horizontal  ones  with  double-entrance  pipes,  as  these  allow  a 
lower  velocity  for  the  same  amount  of  water.  Fig.  112  shows  graphically 


120  CENTRIFUGAL  PUMPING  MACHINERY 

the  velocities  in  a  single-entrance  vertical  pump,  which  can  be  compared 
with  the  ones  in  a  same  size  horizontal  or  vertical  pump  having  double- 
entrance  pipes,  shown  in  Fig.  113,  the  capacities  and  heads  being  the  same. 

NEW  DRY-DOCK  AT  NORFOLK  NAVY  YARD. 

The  new  dry-dock  and  pumping  equipment  at  Norfolk  Navy  Yard  were 
completed  in  1909.  The  pump  house  is  located  at  the  side  of  the  dock  and 
below  the  surface  of  the  ground,  as  shown  in  Fig.  114.  The  dock  is  known 
as  No.  3,  and  the  pump  well  is  complete  with  roof,  galleries,  stairs,  mezza- 
nine floor,  and  ladders  to  the  suction  pit,  to  give  easy  access  to  all  parts. 

Owing  to  the  fact  that  no  anchor  bolts  could  be  secured  to  the  walls  or 
well  bottom,  it  was  necessary  to  carry  the  entire  weight  of  the  pumping 
machinery  on  cross  beams  and  on  trusses  set  into  pockets  in  the  side  walls, 
which  are  clearly  shown  in  Fig.  114.  In  running,  no  vibrations  are  set  up, 
which  proves  the  ability  of  the  structure  to  easily  support  the  entire  weight. 
The  plant  comprises  two  54-inch  double-suction  volute  pumps,  and  two  12- 
inch  drainage  pumps,  including  all  gate  valves,  operating  gears,  and  motors. 
The  main  motors  are  550-horse-power  three-phase  60-cycle  220-volt  in- 
duction motors,  operating  at  a  maximum  speed  of  200  r.p.m.  when  the 
secondary  windings  are  short-circuited.  The  rotors  are  equipped  with 
collector  rings,  and  full  controller  and  resistances  for  starting  under  full- 
load  torque,  with  only  full-load  current,  and  operated  continuously  at  any 
speed  from  75  per  cent  to  full-load  speed.  The  motor  frames  have 
ventilating  openings  allowing  free  circulation  of  air  around  the  windings. 
The  windings  are  covered  with  impregnated  insulation  to  prevent  damage 
from  dampness,  and  have  been  tested  to  stand  an  alternating  electromotive 
force  of  5500  volts  for  one  minute.  The  shafts  are  open-hearth  steel  and  all 
bearings  are  self-oiling  and  self-aligning,  with  ample  surface  to  insure  cool 
running.  The  rise  in  temperature  in  the  motors  does  not  exceed  40°  C. 
with  the  surrounding  air  at  temperature  of  25°  C.  Fig.  82  shows  the  char- 
acteristic curves  of  the  motors. 

The  drainage-pump  motors  are  of  the  constant-speed,  60-cycle  three- 
phase,  220-volt  induction  type,  with  a  normal  output  of  60  horse  power  at  a 
speed  of  514  revolutions,  and  are  provided  with  starting  devices  so  that 
the  motors  can  be  gradually  brought  up  to  full-load  speed  and  full-load 
torque  with  no  more  than  full-load  current.  These  motors  possess  the 
same  details  of  winding,  ventilating  devices,  shafts,  and  bearings  as  the 
others,  and  were  subjected  to  the  same  temperature  tests. 

There  is  also  provided  a  transformer  bank,  consisting  of  three  50-kilo- 
watt  transformers  for  operating  the  capstan  and  valve  motors,  which  are 
of  the  2300-230-volt  oil-insulated  single-phase  type.  These  were  operated 
for  twelve  hours  continuously  with  2300  volts  primary,  at  the  rated  out- 
put in  amperes  and  a  unity  power  factor,  and  heated  to  only  35°  C.  with 


DOCKS 


121 


122  CENTRIFUGAL  PUMPING  MACHINERY 

the  surrounding  air  at  50°  C.  At  the  end  of  the  run  the  load  was  taken  off 
and  the  rise  in  potential  did  not  exceed  40  volts.  An  alternating  electro- 
motive force  of  10,000  volts  was  applied  for  five  minutes  between  primary 
and  secondary,  the  latter  to  the  core,  together  with  an  alternating  electro- 
motive force  of  4000  volts,  momentarily  applied  between  the  low-tension 
winding  and  the  core.  Each  transformer  was  also  operated  for  two  hours 
continuously  with  2300  volts  primary  and  an  output  of  50  per  cent  in  ex- 
cess of  normal. 

There  is  also  provided  a  complete  lighting  transformer  of  the  5-kilo- 
watt  2300-volt,  oil-insulated  single-phase  type,  for  operating  motor-driven 
fan  and  lights. 

A  full  switchboard  of  four  panels  carried  on  suitable  iron  framework, 
two  panels  having  three-pole  single-throw  300-ampere  automatic  oil 
switches  and  250-ampere  ammeter  for  the  550-horse-power  motors  and 
similar  ones  for  the  60-horse-power  motors.  The  third  panel  controls  the 
bank  of  three  50-kilowatt  transformers  and  the  fourth  controls  the  capstan 
and  lighting  as  well  as  the  valve  motors. 

The  gate  valves  for  the  pumps  consist  of  two  60-inch  suction  valves  and 
two  54-inch  discharge  valves,  besides  smaller  valves  on  the  drainage  pumps. 
They  are  of  the  double-gate  type,  with  iror  bodies  flanged  at  both  ends  and 
bronze  fittings  throughout.  Each  valve  is  operated  by  three-phase  60- 
cycle  constant-speed  230-volt  induction  motors,  working  through  cone 
friction  clutches,  fitted  with  an  adjustment,  so  that,  in  case  of  obstruction 
in  the  movement  of  the  gate  valves  or  failure  of  the  limit  stops,  the  friction 
clutches  will  slip  without  overloading  the  motors.  The  arrangement  of 
the  clutches  is  such  that  the  motors  can  be  started  light  and  the  loads 
applied  gradually.  All  of  these  valves  are  arranged  to  be  operated  also  by 
hand  through  extension  stems  and  hand  wheels,  and  are  fitted  with  limit 
stops  and  indicators.  The  drainage  pumps  are  fitted  with  16-inch  valves 
on  the  suction  and  12-inch  valves  on  the  discharge,  with  equipment  similar 
to  that  of  the  large  ones. 

The  main  pumps  are  of  the  double-suction  volute  type,  horizontally 
arranged  on  the  shaft  and  directly  connected  to  the  550-horse-power  motors 
through  special  couplings.  The  motors  and  pumps  are  mounted  on  con- 
tinuous bedplates  for  proper  alignment.  The  pump  casings  are  in  halves 
parted  on  the  horizontal  line  in  order  to  give  easy  access  to  the  internal 
working  parts.  Suitable  manholes  are  provided  on  the  casings  to  facilitate 
internal  inspection  and  cleaning.  The  suction  elbows  are  made  in  halves 
to  allow  the  removal  of  the  impellers  and  shafts  without  disconnecting  the 
couplings  from  the  shafts.  The  bronze  shaft  bearings  are  in  halves  fitted 
to  the  hubs  of  the  suction  elbows,  and  are  provided  with  lubricating  de- 
vices. The  main  casings  are  of  the  volute  type,  with  a  diffusion  throat 
designed  to  give  the  maximum  of  efficiency.  The  casings  are  ribbed  to> 


DOCKS 


123 


rithstand  any  stress  to  which  they  may  be  subjected.  The  impellers  are 
the  inclosed  type,  with  passages  connected  to  the  two  suction  elbows, 
turned  and  polished  all  over  to  minimize  skin  friction,  and  are  balanced  so 
as  to  run  true.  The  shafts  are  of  the  best  nickel  steel,  ground  and  highly 
finished,  and  fitted  with  couplings  at  one  end  to  connect  to  the  motor. 
The  bearings  and  stuffing  boxes  are  composition-bushed  and  provided 
dth  lubricating  devices.  * 

The  drainage  pumps,  smaller  in  size  and  of  vertical  form,  possess  the 
same  details  of  construction. 

The  main  pumps  were  each  required  to  handle   an   average   of  not 
less  than  68,000  gallons  per  minute  when  starting  against  a  static  head 


0          10         20         30         -10         50         60         70         80         90        100 
Per  Cent  Eff. 

Fig.  115.     Performance  Curve  for  Total  Plant  Dry-dock  Work. 

zero  feet  and  ending  with  a  static  head  of  36  feet  through  the  system 
)f  piping.     The  power  delivered  by  the  motor  to  the  impeller  shaft  at 


50-60         70         80         90        100 
Capacity  1000  G.P.M. 


120 


Fig.  116.     Pump  Performance  Curves  for  Dry-dock  Work. 

no  time  could  exceed  550  horse  power.     The  average  actual   capacity 
of  the  pumps  was  76,000  g.p.m.  against  a  total  head  of  36  feet.     The 


124 


CENTRIFUGAL   PUMPING  MACHINERY 


US' 


24'        20'       16'       12'         8' 
Depth  of  AVater  in  Dock 


Fig.  117.     Pump  Performance  Curves  for  Dry-dock  Work. 

dock,  with  a  capacity  of  17,000,000  gallons,  was  emptied  in  two  hours. 
Fig.  115  gives  the  performance  of  the  entire  plant  as  tested  at  the  Navy 
Yard,  Norfolk,  Va.  The  individual  performance  of  these  pumps  is  shown 
in  Fig.  116,  giving  the  characteristic  curves  as  usually  laid  out;  Fig.  117 


24  20  10  12 

Depth  of  Water  in  Dock 


500 


100 


Fig.  118.     Pump  Performance  Curves  for  Dry-dock  Work. 

shows  the  characteristic  curves  based  on  dock  levels;  Fig.  118  shows 
the  average  performance  for  the  whole  unit,  including  motors  and  piping. 
These  curves  show  the  exceedingly  high  efficiency  of  92  per  cent  for  the 


90          86  82          78  74 

Elevation  of  Water  in  Dock 


Fig.  119.     Pump  Performance  Curves  for  Dry-dock  Work. 


DOCKS 


125 


Fig.  120.     Recent  Type  Dock  Pump  at  Tees  Dock,  Middlesbrough,  England. 


Fig.  121.     Another  Type  of  Dock  Pump. 


126 


CENTRIFUGAL  PUMPING  MACHINERY 


pump  alone,  with  allowance  for  frictional  resistance  through  the  piping  and 
valves.     The  efficiency,  including  pipe  and  valve  friction,  reaches  85  to 


Fig.  122.     Plan  of  Installation  at  Tees  Dock,  Middlesbrough,  England. 


Fig.  123.     Plan  of  Installation  at  Tees  Dock,  Middlesbrough,  England. 

88  per  cent  for  a  capacity  of  68,000  g.p.m.  at  about  30-foot  head,  and  re- 
mains practically  constant  up  to  total  head  of  36  feet,  the  capacity  dropping 
to  62,000  gallons. 


DOCKS 


127 


This  installation  is  the  most  comprehensive  and  up  to  date  in  dock  pump- 
ing, and  forms  a  valuable  addition  to  the  Navy,  as  it  provides  for  the  dock- 
ing of  the  new  large  battleships. 

Another  dry-dock  performance  from  the  Portsmouth   Navy  Yard  is 
shown  in  Fig.  119,  showing  an  average  efficiency  of  43.2 
per  cent  for  motors,  pumps,  and  piping,  including  all 
losses. 

Representative  types  of  dock  pumping,  illustrating 
some  later  developments,  are  shown  in  Figs.  120,  121, 
122,  and  123,  the  first  two  showing  30-inch  pumps,  and 
the  latter  two  an  installation  at  Tees  Dock,  Middles- 
brough, England,  consisting  of  48-inch  pumps  with  a 
mean  capacity  of  42,000  g.p.m.  and  34  feet  total  head 
at  295  r.p.m.  These  pumps  gave  a  maximum  efficiency 
as  high  as  85  per  cent  for  the  pumps  themselves,  and 
the  power  of  the  motor  at  the  lowest  head  did  not 
exceed  by  more  than  16  per  cent  that  required  at  the 
point  of  best  efficiency.  Each  pump  is  operated  by  a 
400-horse-power  three-phase  motor. 

Fig.  124  shows  the  usual  arrangement  for  pontoon 
or  floating  dock  pumps,  where  the  pumping  conditions 
vary  from  those  of  graving  docks,  because  the  upper 
surface  is  uncovered  and  the  weight  of  the  ship  is  sup- 
ported by  the  dock.  The  side  sections  of  a  pontoon 
dock  are  small  and  therefore  quickly  emptied.  The 
difference  in  the  level  of  the  water  within  the  dock  and 
without  is  to  be  taken  into  account,  as  the  head  curve 
shows  an  increasing  lift  from  the  start  until  the  floor 
is  uncovered,  then  a  rather  quick  reduction  of  lift, 
which  gradually  increases  until  the  dock  is  emptied. 

Fig.  125  shows  part  of  the  installation  at  the  League 
Island  or  Philadelphia  Navy  Yard,  consisting  of  four 
54-inch  units,  mounted  in  pairs  on  one  continuous 
bedplate.  Fig.  126  shows  a  vertical  installation  for 
a  Japanese  Navy  Yard,  containing  three  48-inch 
pumps. 

Careful  construction  of  the  integral  parts  of  such 
pumps  makes  it  possible  to  obtain  efficiencies  as  high 


Fig.  124.  Arrange- 
ment for  Pontoon 
or  Floating  Dock 
Pump. 

as  90  per  cent  under  the  favorable  conditions  of  dock  work,  handling  a  large 
amount  of  water,  and  demonstrates  that  this  type  is  unquestionably  the 
most  economical  one.  It  is  in  such  installations  that  the  centrifugal 
pump  is  at  its  best,  and  results  can  be  obtained  which  cannot  be  surpassed 
by  any  other  type. 


128  CENTRIFUGAL  PUMPING  MACHINERY 


Fig.  125.     Part  of  Installation  at  Philadelphia  Navy  Yard. 


Fig.  126.     Vertical  Pump  at  Japanese  Navy  Yard. 


CHAPTER  XX. 
CENTRAL  FIRE-STATION  SERVICE. 

IN  fire  service,  centrifugal  turbine  pumps  have  been  a  success  from  the 
beginning.  This  is  largely  due  to  the  fact  that  they  have  been  built  to 
suit  special  requirements,  which  have  been  carefully  studied  out  by  the 
insurance  companies  and  their  board  of  inspection.  Much  of  the  trouble 
found  in  the  installation  of  a  turbine  pump  is  due  to  insufficient  or  incorrect 
knowledge  of  the  exact  requirements  and  pumping  conditions.  Appreci- 
ating this  fact,  the  Associated  Factory  Mutual  Fire  Insurance  Companies 
developed  a  set  of  specifications  covering  the  essential  features  under  which 
many  pumps  have  been  built  which  are  giving  good  results. 

The  essential  characteristics  of  Underwriter  turbine  fire  pumps  are 
ruggedness  and  strength,  liberal  water  passages,  noncorrosive  material 
for  all  working  parts,  ease  in  dismantling,  and  certain  special  features  for 
fire  fighting.  These  are  not  secured  at  any  sacrifice  of  simplicity  and 
reliability,  and  with  the  help  of  such  specifications  the  actual  responsibility 
as  to  design  and  efficiency  lies  with  the  designer  and  manufacturer. 

Each  manufacturer  must  have  the  drawings  and  details  approved  and 
a  sample  of  the  pump  carefully  tested  out  at  the  factory  under  supervision 
of  the  underwriters.  After  this  the  manufacturer  must  agree  that  all  sub- 
sequent pumps  shall  be  equal  to  the  sample  tested,  and  that  no  changes 
in  design  will  be  made  without  the  sanction  and  approval  of  the  board. 

These  pumps  may  be  run  by  a  motor  or  a  steam  turbine;  the  former, 
however,  is  subject  to  the  risk  of  loss  due  to  the  interruption  of  the  electric 
current  and  is  not  considered  as  good  a  fire  risk.  In  order  to  make  this 
satisfactory,  the  source  of  electric  power  should  be  made  as  reliable  as  a 
steam  supply. 

There  are  four  standard  sizes  of  these  pumps  of  the  following  capacities : 
500,  750,  1000,  and  1500  gallons  per  minute.  They  are  made  for  speeds 
suitable  for  the  standard  motors  and  steam  turbines.  The  efficiencies 
required  are  between  50  and  70  per  cent,  depending  upon  the  size,  and  are 
considered  reasonable. 

Each  pump  is  required  to  discharge  a  certain  portion  of  its  full  capacity 
against  a  high  pressure  and  considerably  more  than  its  capacity  against  a 
pressure  of  75  pounds  without  overloading  the  motor  more  than  25  per 
cent.  The  higher  pressures  are  needed  for  high  buildings  or  for  fires  at 
distances  which  require  long  lines  of  hose.  The  pumps  will  give  the  neces- 

129 


130 


CENTRIFUGAL  PUMPING  MACHINERY 


sary  pressures  within  a  certain  range,  with  constant-speed  motors,  but  it  is 
better  to  use  variable-speed  motors. 

The  pumps  should  have  the  supply  under  a  head  for  priming,  as  they 
will  not  pick  up  their  suction  water.  There  is  a  possibility  of  making  an 
automatic  source  of  priming  supply.  In  any  case,  for  fire  purposes  a 
special  priming  tank  of  about  four  times  the  capacity  of  pump  casing  and 
pipes  should  be  installed.  On  long  suction  pipes  and  large  installations 
an  independent  motor-driven  air  or  vacuum  pump  should  be  connected 
to  the  pump  casing  for  exhausting  the  air.  Where  a  reliable  supply  of 
steam  or  compressed  air  is  available,  an  ejector  or  exhauster  may  be  used, 
but  care  should  be  taken  to  have  it  properly  proportioned  for  its  service. 


Fig.  127.     Standard  Underwriter  Type,  a  1000-gallon  Fire  Pump  with  Waterproof 

Motors. 

When  foot  valves  are  used  they  should  be  of  the  multiflap  type,  with  an 
area  of  at  least  150  per  cent  of  the  pipe  area,  and  the  flaps  should  open 
toward  the  sides  of  the  valve  so  as  to  give  an  unobstructed  opening. 
Screens  should  have  a  clear  area  of  200  per  cent  of  the  pipe  area. 

If  check  valves  are  used  on  the  discharge,  they  should  also  have  the 
clapper  or  valve  folding  back  against  the  wall. 

The  motors  should  not  burn  out  if  all  the  streams  are  shut  off  or  when  they 
are  opened  up  to  more  than  the  capacity  of  the  pump,  and  they  should 
easily  stand  a  load  of  25  per  cent  over  that  for  which  the  pump  is  built, 
and  should  be  protected  from  possible  leakages  from  the  pump. 

The  control  affected  by  throttling,  speed  variation,  or  by  both,  should  be 


CENTRAL   FIRE-STATION  SERVICE 


131 


considered.  The  throttling  of  a  pump  should  be  done  in  the  discharge 
valve  and  not  in  the  suction  valve.  Speed  control  is  more  satisfactory 
and  can  be  obtained  with  both  motors  and  steam  turbines.  There  are 
several  classes  of  electric  motors  used  for  fire  pumps,  —  alternating-current 
motors  of  constant-speed  induction  type,  and  direct-current  motors.  In 
the  induction  type  of  motor,  with  the  pump  discharge  valve  closed,  the 
torque  required  is  about  oCTper  cent  of  that  for  full  load.  From  this  it  can 
be  seen  that  the  motors  may  be  readily  started  on  about  65  per  cent  of  the 
regular  voltage  and  brought  up  to  speed  with  not  more  than  double  the  full- 
load  current.  Full  voltage  can  be  obtained  with  a  little  increase  in  current 
and  the  valve  opened  for  the  desired  delivery. 


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100  200  300  400  500  GOO  700  800  9001000   1200 

Gall.  Per  Min. 


1400        1600 


Fig.  128.     Performance  Curve  Underwriter  Pump. 


The  method  of  regulating  shunt-wound  or  direct-current  motors  of 
variable  speed  is  simple  and  well  understood,  and  the  variable-speed  in- 
duction motor  is  similar  to  the  direct-current  armature  in  control. 

Fig.  127  illustrates  a  1000-gallon  fire  pump  with  waterproof  motors  of 
the  standard  Underwriter  type. 

Fig.  128  shows  the  characteristics  for  a  1000-gallon  four-stream  pump, 
designed  to  meet  all  the  requirements,  and  shows  that  on  the  Underwriter 
basis  of  250  gallons  per  stream,  2j-inch  nozzle,  the  pumps  would  give 

4  streams  at  240  feet  head,  or  103  pounds  pressure; 

5  streams  at  185  feet  head,  or  80  pounds  pressure; 

6  streams  at  105  feet  head,  or  45  pounds  pressure; 
3  streams  at  290  feet  head,  or  126  pounds  pressure; 
2  streams  at  310  feet  head,  or  135  pounds  pressure. 


132  CENTRIFUGAL  PUMPING  MACHINERY 

A  maximum  head  of  318  feet,  or  137  pounds  pressure,  is  reached  with  a 
capacity  of  from  300  to  500  gallons  or  one  to  two  streams.  With  this 
range  it  can  be  seen  that  the  centrifugal  fire  pump  can  compete  with  the 
reciprocating  fire  pump. 

The  horse-power  output  of  the  motor  never  reaches  100,  while  the  Under- 
writer requirements  for  this  size  of  pump  allows  for  100  to  107  horse  power. 
The  theoretical  horse  power  required  to  discharge  one  stream  of  250  gal- 
lons under  a  pressure  of  100  pounds  is  14.61,  to  which  must  be  added  the 
lift  of  the  suction  supply  and  the  losses  in  the  pipes,  or  about  1.4  horse 
power,  making  a  total  of  16  theoretical  horse  power  for  each  fire  stream. 

The  following  is  the  record  of  a  test  of  a  1000-gallon  centrifugal  fire  pump 
made  under  the  supervision  of  the  inspectors  of  the  Associated  Factory 
Mutual  Fire  Insurance  Company,  Nov.  29,  1910.  The  pump  was  designed 
for  a  100-horse-power  induction  motor,  but  the  test  was  made  with  a  direct- 
current  shop  motor,  temporarily  connected  to  the  pump.  The  speed  of 
the  motor,  with  no  load,  was  given  on  the  plate  as  1200  r.p.m.  Assuming 
some  falling  off  from  this  speed  under  load,  the  test  was  made  at  as  nearly 
1150  r.p.m.  as  was  practicable.  The  suction  was  taken  from  a  submerged 
tank  through  a  strainer  and  foot  valve ;  the  discharge  was  conducted  to  a 
large  box  with  properly  fitted  baffle  plates  and  a  36-inch  weir.  Two  tests 
were  made  with  three  IJ-inch'and  four  IJ-inch  Underwriter  play  pipes, 
discharging  through  50  feet  of  cotton  rubber-lined  hose  into  the  weir  box, 
and  they  checked  up  reasonably  close  with  the  weir  measurements.  The 
revolutions  were  obtained  by  an  English  tachometer  checked  up  occasion- 
ally by  an  ordinary  speed  indicator.  The  pump  was  run  without  stopping 
for  about  an  hour  and  a  half.  No  difficulty  was  experienced  with  any 
of  the  bearings.  The  behavior  of  the  pump  under  all  loads,  from  no  load 
to  overload,  was  very  satisfactory.  The  use  of  three  stages  in  the  con- 
struction of  this  pump  makes  it  possible  to  secure  a  much  higher  pressure 
when  running  at  half  capacity  than  would  be  possible  were  only  two  stages 
employed. 

The  pump  was  tested  up  to  240  pounds  pressure  and  showed  no  weak- 
ness. Subsequently  the  pump  was  opened  and  the  first  impeller  and  sec- 
tion of  casing  removed. 

The  results  of  the  test  are  given  in  the  accompanying  table.  The 
capacity  of  1000  gallons  per  minute  was  obtained  under  a  total  head  of 
236  feet  at  1150  r.p.m.  and  with  an  efficiency  of  66J  per  cent.  The  motor 
horse-power  output  drops  off  at  the  higher  discharges,  so  that  it  is  not  likely 
that  the  motor  could  be  overloaded.  The  electrical  readings  for  computing 
the  efficiency  of  pump  were  taken  from  Weston  instruments  used  in  the 
regular  testing  work  of  the  shop. 

One  of  the  first  large  cities  to  install  this  type  of  fire  pump  was  Brooklyn, 
N.  Y.,  followed  by  New  York  and  San  Francisco.  Philadelphia,  Winnipeg, 


CENTRAL   FIRE-STATION   SERVICE 


133 


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134  CENTRIFUGAL  PUMPING  MACHINERY 

and  the  New  York  branch  at  Coney  Island  installed  gas-engine-driven  tri- 
plex reciprocating  power  pumps.     A  brief  description  of  the  first  high- 
pressure  fire-service  turbine  pumping  plant  installed  at  Brooklyn,  N.  Y. 
will  be  of  interest. 

This  consists  of  two  separate  stations,  one  situated  at  corner  of  Wil- 
loughby  Avenue  and  St.  Edwards  Street,  and  comprises  three  units.     The 
other  or  main  station  is  at  the  northeast  corner  of  Furman  and  Joralemon 
Streets  and  has  eight  units.     Unlike  the   New  York  high-pressure  fire- 
service  stations,  both  these  plants  can  secure  water  from  either  a  salt-j 
water  or  a  fresh-water  supply,  requiring  a  more  careful  design  in  order  toj 
give  the  same  results  when  working  with  a  suction  head  of  50  pounds  pressure 
and  when  lifting  the  suction  water  from  a  supply  20  feet  below  the  pumps. 
The  salt-water  supply  comes  from  the  East  River,  and  the  fresh-water  from 
a  20-inch  main.     This  arrangement  is  of  great  importance,  as  it  insures  &j 
fire  stream  even  though  the  city  main  that  supplies  the  suction  water 
should  break.     It  is  usually  required  that  a  pumping  station  relying  upon 
a  salt-water  supply  should  be  fitted  up  so  that  equally  good  results  can  be 
obtained  with  salt  water  or  fresh. 

In  installing  the  high-pressure  fire-service  stations  in  the  Borough  of 
Brooklyn,  equal  efficiency  was  demanded  under  both  conditions. 

The  main  station  at  the  foot  of  Joralemon  Street  consists  of  eight  units, 
as  shown  in  Fig.  129.     The  relief  station  at  Willoughby  Street  is  shown  in] 
Fig.    130.     The   pumping   units   are   all   identical,   and   interchangeable. 
Fig.  131  shows  one  of  the  units  complete.     It  consists  of  a  six-stage  turbine 
pump  directly  connected  to  the  motor.     The  pumps  are  of  the  horizontal 
type,  each  capable  of  delivering  3000  gallons  per  minute,  when  operating 
at  735  revolutions  and  delivering  against  a  pressure  of  300  pounds  per; 
square  inch,  with  a  suction  lift  of  20  feet,  or  under  a  supply  pressure  of  50 
pounds  per  square  inch  or  less  when  taking  supply  from  the  20-inch  main. 
The  size  of  the  suction  openings  for  each  pump  is  12  inches  and  the  dis-j 
charge  10  inches.     These  pumps  so  operate  that  the  pressure  system  can  be" 
regulated  between  100  to  300  pounds,  by  increments  of  50  pounds,  the 
speed  remaining  constant,  the  regulation  being  effected  by  the  design  of 
impeller  and  by  special  regulating  valves.     The  brake  horse  power  of  the 
motor  when  operating  at  any  lower  pressure  than  300  pounds  does  not 
exceed  that  at  300  pounds,  thereby  obviating  overloading  the  motor  when 
the  capacity  increases  and  the  pressure  drops. 

Each  pump  is  directly  connected  to  a  three-phase  25-cycle  6000-6300- volt 
800-b.h.p.  motor  of  the  induction  type,  and  fitted  with  all  the  necessary 
starting  and  controlling  devices.  The  full-load  efficiency  of  these  motors 
is  95  per  cent,  the  power  factor  96,  and  the  slip  2  per  cent.  At  three- 
quarters  load,  the  motor  efficiency  is  95  per  cent  and  the  power  factor  95J 
per  cent. 


CENTRAL   FIRE-STATION   SERVICE 


135 


136 


CENTRIFUGAL   PUMPING   MACHINERY 


CENTRAL   FIRE-STATION   SERVICE 


137 


.s 

ex 


138 


CENTRIFUGAL  PUMPING  MACHINERY 


The  results  of  the  tests  given  in  Figs.  132, 133,  and  134  show  an  efficiency 
on  salt  water  of  75  per  cent  and  on  fresh  water  of  76  per  cent  for  the  pumps, 


425 
400- 
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Gallons  per  Minute 

Fig.  132.     Performance  Curve  for  Brooklyn  Fire  Pump. 

and  an  over-all  efficiency  for  motor  and  pump  of  72.2  per  cent.  The  water 
was  measured  by  a  Venturi  meter  in  the  mains.  The  capacity  of  the  pumps 
at  various  pressures  is  shown  in  the  curves,  as  well  as  the  horse  power  re- 
quired for  operation,  and  the  characteristics  controlling  them. 


Gallons  pei  Minute 
Fig.  133.     Performance  Curve  for  Brooklyn  Fire  Pump. 

When  operating  with  salt  water  the  pumps  are  primed  by  a  motor-driven 
vacuum  pump  operating  automatically.  It  was  found  that  with  either 
salt  or  fresh  water  the  pumps  can  be  started  and  brought  to  full  speed  in 
less  than  45  seconds,  without  having  the  starting  current  exceed  150  per 
cent  of  the  full-load  current  of  the  motors.  An  additional  time  of  10 
seconds  was  required  to  bring  the  pump  up  to  a  pressure  of  100  pounds. 
Therefore,  a  unit  can  be  started  from  complete  rest  and  brought  up  to 


CENTRAL   FIRE-STATION   SERVICE 


139 


speed  under  a  water  pressure  of  100  pounds  in  less  than  one  minute.  The 
pressure  can  be  varied  from  300  pounds  down  without  overloading  the 
motor,  an  extremely  severe  condition. 

It  is  interesting  to  note  the  results  obtained  by  gas-engine-driven  power 
pumps  for  similar  purposes,  notably  those  at  Philadelphia  and  Winnipeg. 
The  mechanical  efficiency  at  Winnipeg  is  from  77  per  cent  to  82  per  cent 
on  pumps  of  1800  gallons  per  minute  capacity,  and  the  pressure  300  pounds, 
giving  a  B.t.u.  efficiency  per  indicated  horse  power  of  8600,  and  on  water 
horse  power  of  11,774.  The  Philadelphia  plant  gave  11,160  B.t.u.  per  indi- 
cated horse  power  and  the  Coney  Island  plant  gave  12,682  B.t.u.  per  water 
horse  power.  Such  gas-engine-driven  fire  pumping  equipments  operated 
on  producer  gas  will  give  from  180  millions  to  210  millions  duty  in  foot 


.3^800 


Gallons  per  Min. 

Fig.  134.     Performance  Curve  for  Brooklyn  Fire  Pump. 

pounds  of  work  done  per  one  million  B.t.u.  consumed,  which  is  a  very  low 
cost  for  operating,  but  as  they  are  not  operated  continuously  this  is  not  of 
great  consideration.  They  have  the  advantage  of  low  cost  of  installation 
and  operation,  but  these  are  offset  by  the  flexibility  of  turbine-pump  instal- 
lations in  handling  larger  capacities  at  any  desired  pressure. 

Electrical  or  turbine-driven  centrifugal-pump  fire  stations  are  considered 
more  reliable  than  gas  installations,  and  the  question  of  economy  in  opera- 
tion is  secondary,  because  the  total  cost  of  pumping  is  only  a  small  portion 
of  the  running  expenses. 

The  capacities  of  the  various  stations  built  are  as  follows:  Brooklyn, 
24,000  gallons  per  minute  at  300  pounds  pressure  and  40,000  gallons  at  150 
pounds  pressure;  New  York,  30,000  gallons;  Coney  Island,  4500  gallons; 
Philadelphia,  9000  gallons;  San  Francisco,  20,000  gallons;  Winnipeg,  9100 
gallons.  From  this  it  will  be  seen  that  the  gas  pumping  engines  are  of 
considerably  smaller  capacities.  The  requirements  in  all  such  stations 
should  be  for  a  definite  capacity  and  head,  for  although  it  might  seem  that 
an  excess  capacity  is  good,  it  usually  overloads  the  motors,  and  extra  power 
has  to  be  paid  for. 


140  CENTRIFUGAL  PUMPING  MACHINERY 

High-pressure  fire-service  stations  are  becoming  a  necessity  in  this 
country  owing  to  the  high  buildings  and  the  larger  volumes  of  water  re- 
quired in  fighting  fires.  A  pressure  of  300  pounds  is  usually  needed  to 
overcome  frictional  losses  and  supply  2J-  to  3-inch  nozzles  at  a  pressure  of 
100  pounds.  Portable  fire  engines  to  handle  such  conditions  are  entirely 
out  of  the  question,  and  the  only  practical  means  for  meeting  the  situation 
is,  the  permanent  high-pressure  station. 

Fire  Engines.  —  The  application  of  turbine  pumps  to  automobile  fire 
engines  is  an  interesting  development.  These  pumps  are  made  in  stages 
of  one,  two  or  three,  suitable  for  the  speed  of  the  automobile  engine,  and 
the  capacity  varies  from  250  gallons  to  1000  gallons  for  pressures  of  220 
pounds  down  to  120,  when  delivering  water  through  1000  feet  or  500 
feet  of  2j-inch  hose  and  IJ-inch  nozzles.  A  usual  requirement  is  that  a 
pump  should  deliver!  700  gallons  per  minute  against  a  pressure  of  120 
pounds  and  300  gallons  against  a  pressure  of  200  pounds. 

The  importance  of  obtaining  the  highest  pressure  is  apparent  on  account 
of  internal  friction  in  the  hose  which  reduces  the  effective  pressure  at  the 
nozzle.  The  regularity  of  flow  and  absence  of  water  hammer  make  this 
the  best  solution  of  a  fire  pump  for  a  portable  fire-fighting  apparatus. 

The  action  of  a  turbine  pump  allows  the  shutting  off  of  any  particular 
hose,  nozzle  or  all,  without  tendency  to  burst  the  hose  as  may  be  the  case 
with  reciprocating  or  rotary  fire  pumps.  The  pump  is  considerably 
lighter  and  takes  less  space,  requiring  no  suction  or  discharge  air  cham- 
bers. The  pumps  are  designed  to  fit  on  the  back  part  of  the  chassis  of  an 
automobile  to  be  driven  by  a  shaft  directly  from  the  engine  or  through 
transmitting  gears. 

The  starting  up  when  a  heavy  suction  lift  is  encountered  is  accomplished 
by  means  of  a  small  rotary  vacuum  pump  operated  by  a  magneto  on  the  shaft 
of  the  engine,  which  exhausts  the  air  in  pump  and  pipes.  When  the  tur- 
bine pump  is  primed,  it  is  thrown  into  action  by  the  clutch  and  develops 
immediately  its  head  pressure  and  the  small  vacuum  pump  is  then  cut 
out.  A  special  by-pass  and  valve  are  attached  to  the  pump  to  allow  the 
water  to  pass  from  discharge  to  suction  in  a  small  stream  so  as  to  prevent 
heating  should  all  nozzles  be  shut  off  and  the  clutch  left  in  while  the 
engine  is  running. 

It  is  therefore  apparent  that  rapid  progress  indeed  has  been  made  in 
adapting  the  turbine  pump  to  an  increased  field,  giving  a  better  and  more 
efficient  fire  engine.  In  large  cities  the  automobile  turbine  fire  engine  is 
usually  designed  to  carry  nothing  but  the  pump.  In  smaller  towns  a 
combination  of  fire  engine,  chemical  and  hose  cart  is  desirable,  as  the  most 
useful  apparatus  for  a  fire  department  at  minimum  upkeep  cost.  The 
economic  advantages  of  such  a  combination  seem  beyond  dispute  and  the 
success  is  already  assured.  The  reliability  is  only  a  question  of  design 


CENTRAL  FIRE-STATION    SERVICE  141 

with  proper  power  of  gasoline  motor,  which  should  be  between  100  and 
125  horse  power  for  the  capacities  stated,  and  well  designed  to  stand  up 
under  continuous  service  without  overheating.  The  tendency  in  the 
large  cities  is  to  supplant  the  fire  engine  in  centers  of  business,  by  high- 
pressure  systems  of  independent  mains  and  hydrants,  such  as  have  been 
described  for  New  York  and  Brooklyn,  but  for  the  suburbs  and  smaller 
cities  and  towns  these  independent  fire  engines  will  be  needed. 

The  Underwriters  have  set  a  standard  for  such  fire  apparatus  that  can 
probably  only  be  reached  by  the  application  of  turbine  pumps  to  the  fire 
engines,  just  as  the  turbine  has  supplanted  the  reciprocating  pump  in 
stations  for  high-pressure  fire  mains  and  on  fireboats. 

Such  automobile  fire  engines  will  weigh  complete  from  9000  to  13,000 
pounds,  and  will  travel  at  the  rate  of  30  to  35  miles  an  hour  over  city 
streets. 


CHAPTER  XXI. 
FIREBOATS  AND  SHIPBOARD  SERVICE. 

FIREBOATS. 

TURBINE  pumps  are  extensively  used  for  fireboat  equipments  where 
reciprocating  flywheel  pumps  were  at  one  time  considered  indispensable. 
The  first  boats  so  equipped  were  the  "  James  Duane  "  and  "  Thomas  Wil- 
lett,"  belonging  to  the  New  York  fire  department;  since  then,  however, 
Chicago,  San  Francisco,  Seattle,  and  Duluth  have  followed  the  example 
set  by  New  York. 


Fig.  135.     Worthington  Steam-turbine  Pump  used  by  New  York  Fire  Department. 

The  New  York  system  of  protecting  its  water  front  adds  considerably  to 
the  shipping  value  of  the  harbor,  and  to  the  value  of  property  along  the 
water  front.  Fig.  135  illustrates  one  of  the  pumps  installed  on  these  boats. 
The  boats  are  duplicates  in  all  respects  and  are  123  feet  long,  27  feet  wide, 
and  14  feet  deep.  Each  boat  contains  two  units  of  two-stage  pumps,  piped 
to  run  in  series  or  parallel,  and  each  boat  is  capable  of  handling  9000  gallons 
per  minute  against  150  pounds  pressure,  or  half  that  capacity  at  300  pounds 
pressure  for  use  in  fighting  land  fires  along  the  water  front  where  the  dis- 
tance is  too  great  to  be  reached  by  direct  streams.  The  pumps  are  driven 

142 


FIREBOATS  AND   SHIPBOARD   SERVICE 


143 


by  600-horse-power  General  Electric  steam  turbines  operating  at  a  speed  of 
1800  revolutions. 

The  following  record  of  the  actual  tests  of  these  boats  gives  the  details 
of  performance. 

On  the  "  Thomas  Willett,"  with  both  pumps  running  at  1800  revolutions 
per  minute,  the  following  readings  were  obtained: 

TABLE  NO.  5.  — "THOMAS  WILLETT." 


Pressure  at  nozzles. 

Pressure  at  pumps. 

G.p.m. 

R.p.ra. 

Size, 
2  in. 

Size, 
2J  in. 

Size, 
3  in. 

Size, 
Sin. 

Size, 
3  in. 

Port. 

Starboard. 

Port. 

Starboard. 

0 

0 
0 
90 

0 

0 

0 
90 

100 
115 

115 
90 
230 

102 

115 
115 
90 

100 
110 
112 

85 

160 
170 

175 
140 
155 

160 
170 

175 
140 
310 

9,035 
9,600 
9,625 
11,460 
4,500 

1800 
1800 
1900 
1900 
1780 

1800 

1800 
1900 
1900 
1790 

On  the  "  James  Duane  "  the  following  tests  were  made: 
TABLE  NO.  6.  — "JAMES  DAUNE  " 


Pressure  at  nozzles. 

Pressure  at  pumps. 

G.p.m. 

R.p.m. 

Size,               Size, 
3  in.                3  in. 

Size, 
3  in. 

Port. 

Starboard. 

Port. 

Starboard. 

73 
75 
40 

85 
90 
170 

73 
60 
55 

85 
90 

150 

"i65  " 
175 
165 

"l50" 

125 
165 
175 
330 

5120 
4920 
6280 
8280 
8480 
3900 

1900 

"i900" 

1800 

1800 
1900 
1800 

0 
53 

85 

87 

.1800 
1900 
1800 

On  both  boats  the  vacuum  varied  from  26  to  27  inches,  and  the  steam 
pressure  from  175  to  190  pounds. 

The  pumps  ran  with  no  vibration,  and  standing  on  the  deck  one  would 
scarcely  know  the  pumps  were  running  except  for  the  water  thrown. 

It  was  found  that  the  speed  for  throwing  full  capacity  could  be  obtained 
in  from  25  to  40  seconds.  The  last  test  on  the  "Thomas  Willett"  consisted 
in  putting  the  two  pumps  in  series  for  300  pounds  pressure. 

Fig.  136  illustrates  the  internal  arrangement  of  these  pumps  and  the  de- 
sign of  impellers  for  high  speeds  and  high  pressures.  Fig.  137  shows 
another  size  of  similar  type.  These  illustrations  show  turbine  pumps 
especially  designed  for  steam-turbine  drives  on  fireboats,  where  the  primary 
requirement  is  instant  operation.  The  use  of  turbine  pumps  eliminates 
vibration  in  the  boat  when  running  full  speed,  and  the  ability  to  come  up 


144 


CENTRIFUGAL  PUMPING  MACHINERY 


to  speed  in  20  to  30  seconds  is  something  never  before  accomplished  in  fire 
pumps.     The  possibility  of  closing  down  all  the  hose  streams  suddenly 

A  Section  A-B-C 


Fig.  136.     Internal  Arrangement  of  Turbine  Pump  used  on  Fireboats. 

without  harm  is  welcome  to  the  operator,  as  in  the  reciprocating  type  of 
pumps  care  must  be  taken  to  slow  down  gradually  or  something  will 
break. 


Fig.  137.     Small  Size  Turbine  Pump  used  on  Fireboats. 

The  boats  have  demonstrated  their  readiness  for  service  and  have 
rendered  valuable  services  in  several  large  water-front  fires.  There  can  be 
no  doubt  that  for  fire  service  the  steam-turbine-driven  turbine  pump  is  far 
superior  to  any  flywheel  reciprocating  pump,  and  that  in  time  it  will 
replace  all  others. 

ON  SHIPBOARD. 

Because  of  its  simplicity  and  compactness  the  centrifugal  pump  has  come 
into  very  general  use  on  shipboard.  Some  of  these  applications  have  been 
given  in  the  previous  chapter  and  under  "  Circulating  Water,"  and  so  need 


FIREBOATS  AND   SHIPBOARD   SERVICE 


145 


no  further  mention  here.  This  type  is  also  used  very  extensively  for  han- 
dling ballast  water  from  the  double  bottoms,  ballast  tanks,  and  water-tight 
compartments,  and  also  in  bridge  work  and  the  sanitary  service.  Wrecking 
boats  are  always  equipped  with  two  centrifugal  pumps  with  about  12-  or 
14-inch  discharge  permanently  fixed  on  board.  They  are  engine-driven 
and  of  either  the  vertical  or  horizontal  type.  The  boilers  are  usually  de- 
signed to  work  with  salt  water  and  to  stand  a  lot  of  abuse. 


Fig.  138.     One  of  the  Latest  Circulating  Pumps  of  the  Birotor  Type  used  on  the 
Battleship  "Arkansas"  of  the  Dreadnought  Class. 

Fig.  138  is  a  photographic  view  of  one  of  the  latest  circulating  pumps  of 
the  birotor  type  used  in  marine  practice  and  now  installed  on  the  battle- 
ship "Arkansas"  of  the  dreadnought  class.  A  steam-turbine-driven  circu- 
lating unit  can  be  arranged  to  operate  under  almost  all  conditions,  such  as 
high  pressure  exhausting  to  atmospheric  or  to  the  condenser,  or  with  low- 
pressure  steam  or  as  a  mixed-flow  turbine,  and  will  furnish  circulating 
water  with  absolute  reliability. 

It  is  to  be  noted  that  the  impellers  work  in  parallel,  having  one  main  inlet 
pipe  and  one  outlet  pipe,  with  a  special  opening  on  the  side  of  the  suction 
chamber  for  pumping  out  the  bilges.  The  pump  casing  is  split  horizontally 


146  CENTRIFUGAL  PUMPING  MACHINERY 

in  order  to  facilitate  repairs  aboard  ship,  where  space  is  of  paramount 
importance. 

The  compactness  of  the  design  lends  itself  readily  to  marine  service,  and 
in  this  instance  the  total  length  is  only  1 1  feet  for  a  unit  having  a  capacity 
of  27,000  gallons  with  a  maximum  head  of  about  37  feet,  including  all  losses 
through  tubes,  sea-cock  piping,  and  valves.  The  particular  feature  in  such 
a  pump  is  the  high  water  speed,  which  makes  it  possible  to  design  a  compact 
pump  and  at  the  same  time  to  obtain  the  reasonably  good  efficiency  of  66 
per  cent. 

"The  minimum  requirements  are  15,000  gallons  per  minute  against  a 
head  of  25  feet,  of  which  15  feet  are  suction  lift,  when  working  on  the  bilges. 
For  this  service  there  is  provided  a  side  suction  opening.  The  main  suction 
openings,  through  which  the  circulating  water  for  the  condensers  is  sup- 
plied, consists  of  two  separate  connections  with  pipes  to  the  sea  cocks." 


CHAPTER  XXII. 

SPECIAL  HIGH-SPEED  INSTALLATIONS. 

THE  high  rotative  speed  of  the  steam  turbine,  and  the  convenience  of 
operating  it  direct-connected  to  a  centrifugal  pump,  have  presented  an 
interesting  problem  to  the  pump  designer.  Where  the  quantity  of  water 
is  not  large  a  small  impeller  may  be  safely  used  if  constructed  of  a  material 


Fig.  139.     Novel  Design  of  Turbine-driven  Pump. 

which  will  not  be  cut  by  the  water.  Fig.  139  illustrates  a  rather  novel  de- 
sign of  turbine-driven  pump  with  a  normal  capacity  of  2300  g.p.m.  to  3300 
g.p.m.  and  a  minimum  capacity  of  900  g.p.m.  for  elevator  service  against  a 
head  of  150  pounds  pressure.  The  shaft  is  carried  in  two  spherical  bearings, 
one  near  the  steam  turbine  and  the  other  on  the  outer  end  of  the  pump,  to 
allow  the  shaft  to  take  a  true  position.  There  is  no  thrust  in  either  the 
turbine  or  the  pump,  but  for  safety  a  special  bearing  has  been  arranged  at 
one  end.  The  turbine  and  pump  are  secured  to  a  common  base  plate,  and 
the  unit  is  entirely  contained. 

Pumps  of  this  design  may  be  made  with  single  impellers  up  to  300  feet 

147 


148 


CENTRIFUGAL   PUMPING   MACHINERY 


head,  without  the  need  of  diffusion  vanes.     Efficiencies  of  these  pumps 
vary  from  75  to  80  per  cent. 

Fig.  140  illustrates  a  14-inch  two-stage  high-speed  steam-turbine- 
driven  pump,  arranged  to  run  in  series  or  in  parallel.  It  was  designed  for 
11,000  gallons  per  minute  at  250  feet  head,  or  5500  gallons  at  500  feet  head, 
and  is  driven  by  a  1000-horse-power  turbine  running  at  1600  r.p.m.  The 
test  of  this  unit  is  given  in  the  following  table: 

TABLE  NO.  7.  — TEST  OF  1000-H.P.  TURBINE  DIRECT-CONNECTED  TO 
TWO   14-INCH  WORTHINGTON   TURBINE  PUMPS. 

QUEMAHONING  DAM,  June  7,  1910. 


Series. 

Parallel. 

Number  of  test. 

1 

2 

3 

4 

5 

6 

Steam  Turbine. 

R.p.m.  turbine  

1,631 

1,631 

1,617 

1,603 

1,620 

1,620 

Steam  pressure  abs  

121.6 

122.3 

112.5 

113.6 

124.0 

126 

Quality  steam,  per  cent. 

96.5 

96.4 

97.1 

96.4 

97.4 

97.5 

Inches  vacuum  

25 

25 

25 

25 

25 

25 

Barometer,  inches  

28.5 

28.5 

28.5 

28.485 

28.485 

28.5 

Vacuum  ref.  to  29.92.... 

26.98 

26.98 

27.04 

26.72 

27.24 

27.67 

Temperature  exhaust  .  .  . 

144 

144 

144 

116.6 

110 

111 

Hot  well   deg  F 

58 

58 

57 

58 

59 

60 

Air-pump  suet.,  deg.  F.. 

50 

60.6 

61.5 

67 

68 

71 

Cooling  HaOent.,  deg.  F. 

56 

56 

56 

57.6 

57.5 

58.5 

Cooling  H2O,   leaving, 

deg.  F  

66.5 

69 

64 

58-64.6 

60-76.8 

67-85 

R  p  m   air  pump 

80.3 

81.2 

80.75 

85.3 

80.25 

83.5 

Pump. 

Lift  in  feet  

15 

14.65 

15.6 

16.15 

13.2 

13.0 

First  stage,  Ibs.  press.  .  . 

104.8 

123 

55.5 

100.3 

135.5 

136.3 

First  stage,  ft.  press..  .  . 

242.0 

284 

128 

232.0 

312.0 

314 

Second  stage,  Ibs.  press. 

219.8 

257 

149.2 

101.0 

137.2 

138.0 

Second  stage,  ft.  press.. 

507.0 

294 

342.0 

233.3 

317.0 

319.0 

Dist.  between  gauge,  ft. 

3.0 

3.0 

3.0 

3.0 

3.0 

3.0 

Total  head,  ft  

525 

611.5 

360 

251.8 

330.7 

332.5 

R.p.m.  pump  

1,631 

1,631 

1,617 

1,603 

1,620 

1,620 

Weir  reading,  ft  

.594 

.522 

.652 

.947 

.388 

.280 

Capacity,  gal.  per  min.  . 

5,430 

4,480 

6,180 

11,020 

2,880 

1,790 

Water  horse  power  

722 

691.5 

563 

654 

241 

151 

Condensed  steam  

20,320 

16,256 

12,192 

12,192 

12,192 

4064 

Dry  condensed  steam  .  .  . 

19,600 

15,670 

11,840 

11,720 

11,750 

3960 

Dry  steam  per  hour  .... 

20,400 

19,620 

16,430 

19,450 

13,240 

8150 

Dry  steam  per  hour, 

water  horse  power  . 

28.3 

28.4 

29.3 

29.8 

55 

54 

Francis  formulae  used  for  weir  calculations. 


Where  large  quantities  of  water  are  needed,  an  ordinary  style  of  pump 
cannot  be  used,  since  the  large  diameter  of  the  impeller  gives  a  prohibitive 


Fig.  140.     14-inch  Two-stage  High-speed  Steam-turbine-driven  Pump. 


Sec 


Fig.  141.     24-inch  Trirotor  Pump  Designed  for  Circulating  Water  for  a  Condenser. 


Fig.  142.     Pump  Designed  for  Circulating  Water  for  a  Condenser,  Fitted  with  Diffusion 

Vanes.  (149) 


150 


CENTRIFUGAL  PUMPING  MACHINERY 


peripheral  velocity.  For  this  service  a  multirotor  pump  has  been  designed 
which  bears  the  same  relation  to  the  capacity  of  the  pump  at  high  speed 
that  the  multistage  design  does  to  the  head. 


16  Discharge, 
23^  "Flange, 
16-  l"  Bolts  on 
lOVRadius. 


Fig.  143.     Vertical  Type  of  Birotor. 

Fig.  141  illustrates  a  24-inch  trirotor  pump  designed  for  circulating 
water  for  a  condenser.  It  is  practically  three  small  impellers  working  in 
parallel  so  that  the  proper  linear  velocities  may  be  obtained  with  the  high 


Fig.  144.     A  24-inch  Trirotor  High-speed  Unit. 


SPECIAL  HIGH-SPEED   INSTALLATIONS 


151 


rotative  speed  of  the  steam  turbine.     Fig.  142  illustrates  a  similar  pump 
fitted  with  diffusion  vanes  hi  the  discharge  chamber  and  suitable  for  high 


Fig.  145.    A  Birotor  Type  of  Steam-turbine-driven  Pump  for  Circulating  Water  in 

Power  Stations. 


circulating  heads.     Fig.  143  shows  a  vertical  type  of  birotor  suitable  for 
certain  classes  of  plants  requiring  this  type. 

A  24-inch  trirotor  high-speed  unit  is  shown  in  Fig.  144.     It  is  designed 


Fig.  146.    A  Trirotor  Type  of  Steam-turbine-driven  Pump  for  Circulating  Water  in 

Power  Stations. 


152  CENTRIFUGAL  PUMPING   MACHINERY 

to  handle  20,000,000  gallons  of  muddy  river  water  per  24  hours  at  1400 
r.p.m.  against  a  head  of  130  to  160  feet,  but  at  times  as  low  as  95  feet. 

Figs.  145  and  146  illustrate  a  birotor  and  a  trirotor  type  of  steam-tur- 
bine-driven pump  for  circulating  water  in  power  stations. 

This  application  of  steam-turbine  drive  to  auxiliaries  in  power  stations 
gives  an  economical  installation  and  a  maintenance  cost  less  than  for  any 
other  type. 


CHAPTER  XXIH. 
COMMERCIAL  PUMPS  FOR  GENERAL  INDUSTRIAL  USES. 

IT  is  the  intention  to  here  give  a  somewhat  detailed  description  of  the 
working  parts  of  a  standard  type  of  pump.  Many  industries  permit  the 
use  of  such  a  type  of  pump,  as,  for  instance,  common  contractors'  work, 
irrigation,  and  similar  work.  Fig.  147  shows  one  of  these  commercial 


Fig.  147.     Commercial  Pump. 

pumps  having  an  average  efficiency  and  a  low  first  cost.  For  paper-mill, 
dye-house,  and  sugar-house  work  a  special  type  of  pump  is  required,  in 
which  the  internal  parts  are  easily  accessible  for  cleaning  without  dis- 
mantling the  entire  machine.  This  style  is  shown  in  Fig.  148,  and  has  its 
casing  split  along  the  horizontal  line,  as  shown  in  Fig.  149.  Such  pumps 
are  especially  adapted  for  handling  material  which  leaves  a  deposit  inside 
of  the  pump,  which  must  be  regularly  removed. 

Another  type,  suitable  for  paper  mills,  sugar  houses,  and  gas  houses,  is 
shown  in  Fig.  150,  the  casing  and  impeller  being  designed  to  handle  heavy 
or  thick  liquors.  They  are  suitable  for  pumping  various  kinds  of  liquors 
from  chests  or  storage  tanks  to  digesters  and  engines,  for  returning  water 

153 


154 


CENTRIFUGAL   PUMPING   MACHINERY 


from  the  screens  back  to  the  filters,  for  pumping  clay  water  from  mixers  to 
the  beating  engines,  and  lime  used  in  bleaching  and  chloride  of  lime ;  in  fact 
for  every  conceivable  use  in  a  mill. 


Fig.  148.     Sugar-house  Pump. 

The  material  out  of  which  such  pumps  should  be  made  depends  upon  the 
liquid  to  be  handled.     Most  of  the  trouble  in  such  pumps  can  be  attributed 


Fig.  149.     Sugar-house  Pump  Showing  Method  of  Examination. 

to  lack  of  knowledge  of  the  liquid  with  which  the  pump  is  to  be  used,  and 
if  the  conditions  are  known  a  pump  can  be  made  suitable  to  the  particular 
service. 


Fig.  150.     Paper-mill  Pump. 


Fig.  151.     Vertical  Pump  Showing  Parts. 


(155) 


156 


CENTRIFUGAL  PUMPING  MACHINERY 


Where  a  vertical  arrangement  is  necessary  pumps  like  that  shown  in  Fig. 
151  may  be  used.  A  similar  type  for  larger  sizes  and  higher  heads  is  shown 
in  Fig.  152.  In  many  modern  buildings  having  deep  basements,  where 
water  can  accumulate,  a  type  of  sump  pump,  illustrated  in  Fig.  153,  is  used. 
This  may  be  arranged  to  operate  automatically,  and  as  the  pump  is  noiseless 
it  is  well  adapted  to  buildings,  apartment  houses,  and  hotels.  In  small 


Fig.  152.     Vertical  Turbine  Pump. 

units  the  turbine  pump  driven  by  electric  motors  with  automatic  float 
switches  is  being  largely  used  in  buildings  for  tank  service. 

To  give  the  reader  a  fair  idea  of  how  the  details  of  a  modern  centrifugal 
pump  look,  the  following  illustrations  have  been  inserted.  Fig.  154  shows 
the  pump  itself  and  the  equipment  usually  needed.  The  casing  may  be  of 
cast  iron,  cast  steel,  or  bronze,  according  to  nature  of  material  to  be  handled. 
The  inlet  and  outlet  are  shown  in  their  normal  position,  but  they  are  some- 
times placed  differently.  Fig.  155  shows  the  impellers,  which  are  also  made 
of  a  material  consistent  with  the  properties  of  the  liquid  to  be  pumped. 
These  impellers  are  balanced  lor  high  rotative  speeds  and  may  have  either 


COMMERCIAL   PUMPS 


157 


r-v 


Fig.  153.    Vertical  House  Sump  Pump. 

a  single  side  suction  or  a  double  suction  inlet.  They  are  designed  accord- 
ing to  mathematical  formulas  for  each  specific  duty.  The  hubs  or  flanges 
shown  are  for  the  purpose  of  preventing  leakage  and  for  facilitating  bal- 


Fig.  154.     Horizontal  Turbine  Pump,  Vertical  Split  Casing. 


158 


CENTRIFUGAL  PUMPING  MACHINERY 


anting.     In  the  spaces  between  the  sides  of  the  impellers  and  the  casing 
walls  there  is  an  area  which  would  be  unbalanced,  and,  in  order  to  overcome 


Fig.  155.     Arrangement  of  Impellers  on  a  Multistage  Pump. 

this,  hubs  are  provided  of  equal  diameters  on  each  side,  having  a  running 
fit  of  not  over  jo'W  of  an  inch  and  having  an  end  play  of  TV  inch  to  allow 


Fig.  156.     Thrust  Bearing. 

the  impeller  to  float  to  its  proper  position.  Holes  are  drilled  in  the  im- 
pellers at  the  middle,  inside  of  these  hubs,  allowing  water  under  the  pressure 
to  act  upon  the  area  within  the  rings,  thus  giving  balance  as  near  as  is  prac- 


COMMERCIAL  PUMPS 


159 


ticable.     This  arrangement  is  followed  for  each  impeller,  so  that  there  is  a 
minimum  thrust,  but  to  compensate  for  any  possible  thrust,  whether  due  to 


Fig.  157.     Diffusion  Vanes. 

suction  lift  or  to  the  wear  that  will  take  place  in  time,  a  self-oiling  thrust 
bearing  is  supplied  at  the  outer  end  as  shown  in  Fig.  156.     The  additional 


Fig.  158.     Complete  Diffusion. 

purpose  of  the  thrust  bearing  is  to  provide  a  lateral  alignment  between 
impellers  and  diffusion  vanes.     The  main  bearing  is  integral  with  the 


160  CENTRIFUGAL  PUMPING   MACHINERY 

stuffing  box  and  secured  to  the  casing  by  a  recessed  fitting.  The  stuffing 
box  is  sealed  and  of  the  lantern  gland  design,  with  soft  packing  each  side 
of  the  gland.  Fig.  157  illustrates  the  diffusion  vanes,  through  which  the 
water  is  guided  in  mathematically  calculated  passages,  transforming  the 
velocity  head  into  pressure  with  the  least  possible  losses.  The  material  of 
these  rings  also  depends  upon  the  liquid.  There  are  two  rings,  one  with 
vanes  and  one  without.  Together  they  form  the  complete  diffusion 
chamber  shown  in  Fig.  158. 


PART   IV.— PRIME   MOVERS   FOR   DRIVING 
CENTRIFUGAL   PUMPS. 


CHAPTER   XXIV. 
ELECTRIC   MOTORS. 

THE  great  majority  of  the  centrifugal  pumps  in  service  are  motor-driven. 
In  many  cases  the  importance  of  the  relationship  between  the  pump  and 
the  motor  has  been  too  little  appreciated.  In  order  to  attain  the  best  re- 
sults, the  engineer  must  fully  understand  the  peculiarities  of  the  pump, 
the  variations  in  load  during  operations,  the  characteristics  of  electric 
power,  and  of  the  plant  as  a  whole.  The  exact  capacity  and  rating  of  the 
motor  are  of  great  importance,  especially  where  power  is  bought  on  the 
motor  rating.  If  the  motor  is  too  small,  it  will  be  constantly  overloaded, 
and  if  it  is  too  large,  the  customer  pays  for  power  not  used.  Furthermore, 
the  power  required  at  rated  speed  and  head  should  not  in  itself  determine 
the  size  of  the  motor.  Maximum  conditions  should  also  be  taken  into 
account. 

Having  determined  the  size  of  motor  required,  it  is  of  utmost  importance 
that  the  proper  style  be  selected  for  the  limits  of  current  available.  It  is 
in  the  selection  of  the  class  or  style  of  motor  that  an  understanding  of  the 
characteristics  of  the  pump  is  absolutely  essential,  as  it  is  necessary  to  select 
a  motor  which  will  take  care  of  the  varying  loads  and  meet  the  different 
starting  conditions.  Where  the  head  must  vary,  this  may  be  accomplished 
by  changing  the  speed,  and  a  motor  must  be  selected  which  permits  speed 
regulation.  The  designer  of  the  pump  must,  therefore,  carefully  consider 
the  nature  of  his  motor  when  laying  out  the  characteristics  of  his  impeller. 
On  the  other  hand,  the  electrical  engineer  should  design  his  motor  to  suit 
the  characteristics  of  the  pump. 

A  pump  should  be  designed  for  an  average  head  at  maximum  efficiency 
and  maximum  head  at  average  efficiency,  and  motor  and  pump  should  be 
so  proportioned  that  under  the  varying  conditions  the  motor  will  not  be 
overloaded  except  within  a  certain  range,  and  will  allow  increased  capacity 
for  lower  heads  with  nearly  constant  power.  The  pump  should  be  designed 
for  a  wider  range  than  is  usual  in  order  to  give  the  motor  a  restricted  output 
and  to  prevent  excessive  overloads  under  abnormally  low  heads.  All  pump 
characteristics  should  have  a  flat  curve  over  a  considerable  range,  making 

161 


162  CENTRIFUGAL  PUMPING  MACHINERY 

the  horse  power  nearly  constant  between  the  limits  of  the  working  condi- 
tions. This  is  especially  necessary  with  induction  motors,  as  both  the 
power  factor  and  the  efficiency  are  reduced  at  light  loads. 

The  problem  of  starting  torque  is  easily  solved  with  direct-current  motors, 
provided  their  size  is  properly  selected  to  correspond  with  the  capacity 
and  head  of  the  pump  under  all  conditions  to  be  met.  The  majority, 
however,  of  pump  installations  are  of  the  alternating-current  type.  An 
induction  motor  of  proper  capacity  will  start  a  pump  successfully,  pro- 
vided the  initial  rush  of  current  is  tolerated.  A  synchronous  motor  may 
fail  to  start  the  pump,  as  the  torque  of  this  type  of  motor  is  usually  small, 
and  increases  only  as  synchronism  is  reached;  while  the  torque  of  the  pump 
increases  with  the  square  of  the  velocity,  thus  producing  a  critical  point 
in  starting.  Several  remedies  have  been  suggested,  chiefly  to  start  the 
pump  empty  and  prime  after  full  speed  has  been  reached,  also  to  start  with 
a  smaller  independent  motor  to  build  up  to  speed.  The  load  on  the  motor 
may  be  relieved  in  a  primed  pump  by  closing  the  discharge  valve  and  allow- 
ing the  pump  to  deliver  water  through  a  by-pass  back  to  the  suction  under 
a  lower  head,  removing  a  portion  of  the  load  from  the  motor.  As  the  start- 
ing torque  in  a  synchronous  motor  is  due  to  eddy  currents  and  hysteresis, 
more  to  the  former  than  to  the  latter,  it  follows  that  every  means  of 
increasing  the  eddy  currents  will  help  the  starting  of  the  pump.  The 
synchronous  motor  will  hardly  equal  the  induction  motor  for  starting 
turbine  pumps,  as  it  has  a  large  air  gap  and  leakage  and  an  uneven  distri- 
bution of  the  secondary  winding,  which  produces  a  lesser  torque  than  can 
be  found  in  an  induction  motor. 

Another  condition  to  be  met  with  is  the  change  of  pressure  by  speed 
variation.  In  multipolar  alternating-current  motors  this  can  be  accom- 
plished by  using  a  different  number  of  poles  for  the  various  speeds.  The 
motor  builder  should  know  how  the  load  on  the  pump  varies  with  varying 
head  and  constant  speed.  Variation  in  the  speed  of  direct-current  motors 
is  usually  accomplished  by  resistance  in  the  field,  but  this  has  a  limited 
application  and  may  prove  objectionable  for  commutation,  as  the  fields 
distort  at  the  high  speeds.  The  interpole  variable-speed  motor  with  a 
rheostat  is  better,  although  more  expensive.  In  any  case,  the  motor  should 
be  large  enough  so  that  the  field  rheostat  cannot  cut  in  too  much  resistance 
and  overload  the  motor.  The  speed  of  the  pumps  should  be  properly 
selected  under  such  conditions  to  avoid  endless  trouble  due  to  the  different 
speeds  at  different  heads. 

One  must  consider  that  in  such  combinations  an  increase  of  10  per  cent 
in  speed  may  increase  the  power  50  per  cent  and  cause  a  great  increase  in 
the  armature  losses.  This  is  pointed  out  to  show  the  great  effect  speed 
variation  may  have  in  an  installation  with  a  variable-speed  motor.  Speed 
is  one  of  the  most  sensitive  characteristics,  and  any  unnecessary  variation 


ELECTRIC  MOTORS  163 

is  a  great  absorber  of  power.  In  a  synchronous  installation  the  speed  is 
fixed  by  the  prime  mover  and  the  motor  is  selected  for  maximum  load. 
There  are  various  methods  for  effecting  the  speed  regulation.  For  example, 
a  two-speed  induction  motor  can  operate  at  normal  horse  power  on  eight 
poles  and  at  double  the  horse  power  on  four  poles.  This  is  accomplished 
by  primary  windings  connected  to  consecutive  poles.  As  a  four-pole  motor 
the  connections  are  in  parallel,  and  for  eight-pole  in  series.  These  motors 
can  be  designed  and  built  for  three  speeds,  operating  on  eight,  ten,  and 
sixteen  poles,  by  having  a  primary  with  two  sets  of  windings,  one  to  eight- 
pole  and  sixteen-pole  connections,  and  the  other  for  ten-pole  operations. 
This  method  of  regulation  is  entirely  practical  and  can  be  applied  in  con- 
nection with  turbine  pumps. 

In  vertical  installations  the  motor  should  carry  the  weight  of  all 
revolving  parts  such  as  armature,  impellers,  and  shafting,  in  order  to 
obviate  any  vibration  which  may  be  set  up.  The  pumps  should  be  placed 
on  good  foundations.  If  this  is  not  done  the  vibrations  set  up  will  ruin  the 
motor  or  the  pump. 

Too  much  stress  cannot  be  laid  upon  the  fact  that  the  pump  and  motor 
designers  must  work  together  and  consider  each  other's  problems  more  than 
has  been  done  in  the  past;  for  the  pump  and  motor  must  be  considered 
together,  just  as  a  steam  end  in  a  direct-acting  pump  must  be  designed  in 
connection  with  the  pump  end. 


CHAPTER  XXV. 

« 

STEAM  ENGINES  AND  MISCELLANEOUS. 

SINCE  the  development  of  high-speed  steam  engines,  more  engine-driven 
centrifugal  pumps  have  been  installed.     This  combination  has  met  with 


C£NTtWtj6A"  **" 


Fig.  159.     Vertical  Cross  Compound  Sewage  Pump. 
164 


STEAM  ENGINES  AND  MISCELLANEOUS  165 

favor  in  small  plants  where  the  engineer  in  charge  is  better  acquainted  with 
the  workings  of  a  steam  engine  than  with  a  motor  or  steam  turbine.  The 
speed  of  an  engine-driven  pump  is  necessarily  limited.  It  rarely  exceeds 
800  r.p.m.,  and  generally  is  in  the  neighborhood  of  600  r.p.m.  Such  a  unit 
cannot  be  used,  therefore,  for  all  kinds  of  service,  and  must  be  started 
slowly  to  give  the  engine  sufficient  tune  to  warm  up. 

Gas  and  oil  engines  are  also  extensively  used  direct-connected  to  cen- 
trifugal pumps.  One  of  the  largest  fields  for  this  type  is  in  contractors' 
work  for  emptying  excavations,  sumps,  and  ditches.  Gasoline  and  kero- 
sene-oil engines  are  used  almost  exclusively  for  this  work,  as  they  are  con- 
venient, portable,  and  need  very  little  attention.  Economy  and  efficiency 
are  secondary  considerations. 

Centrifugal  pumps  for  almost  any  kind  of  service  may  be  driven  by  belts 
or  silent  chains  if  shaft  power  is  available,  as  we  need  only  consider  the 
speed  of  the  shafting  and  the  ratio  of  the  driver  and  driven  pulleys  to  obtain 
the  necessary  speed. 

Where  there  is  an  abundance  of  water  under  a  few  feet  of  head  and  a  lesser 
quantity  is  desired  at  greater  head,  a  combination  centrifugal  pump  and 
water  wheel  has  been  found  to  be  very  economical.  The  water  wheel  or 
water  turbine  is  direct-connected  to  a  centrifugal  pump  which  distributes 
the  water  under  the  desired  head  through  a  line  of  piping.  This  combina- 
tion may  also  be  used  where  there  is  an  available  water  power  for  the  water 
wheel,  and  a  separate  supply  of  pure  water  or  some  other  liquid  to  be 
pumped. 

Fig.  159  shows  different  types  of  engine-driven  pumps,  the  former  a 
compound  engine  with  pump  between,  the  latter  the  usual  arrangement  of 
single  engine  and  pump.  These  types  are  commonly  used  for  supplying 
circulating  water  for  condenser,  sewerage  work,  and  similar  duties. 


CHAPTER  XXVI. 

STEAM  TURBINES. 

CENTRIFUGAL  pumps,  driven  by  steam  turbines,  are  being  extensively 
used  for  hot-well,  boiler-feed,  and  circulating  pumps.  These  combinations 
are  efficient,  occupy  small  space,  and  require  minimum  attendance.  In 
the  design  of  such  a  unit  the  chief  problem  is  to  construct  turbines  of  20  to 
500  horse  power  which  can  operate  at  a  speed  suited  to  the  pump.  A 
compromise  must  be  made  between  the  ideal  speeds  of  the  pump  and 
the  turbine. 

In  large  power  plants,  where  steam  can  be  used  for  heating  feed  water, 
the  efficiency  of  an  auxiliary  prime  mover,  like  a  steam  turbine,  is  of  sec- 
ondary importance,  as  low  cost  of  operation  will  offset  increased  steam 
consumption.  The  rating  of  steam  turbines  in  connection  with  pumps 
should  always  be  on  the  maximum  load,  as  otherwise  the  expected  efficiency 
will  not  be  obtained. 

There  are  four  principal  types  of  steam  turbines :  • 

First.  De  Laval;  an  impulse  turbine  in  which  the  steam  is  completely 
expanded  in  a  single  set  of  nozzles  and  all  the  kinetic  energy  is  given  up 
to  a  single  row  of  blades. 

Second.  Parsons ;  impulse-reaction,  where  the  energy  of  reaction  of  an 
expansion  in  the  moving  blades  is  added  to  the  impulse  of  the  steam  as 
received  from  the  fixed  nozzles. 

Third.  Zoelly  and  Rateau;  impulse  turbine  having  a  series  of  partial 
expansions,  the  energy  of  each  expansion  being  absorbed  in  a  single  row 
of  moving  blades. 

Fourth.  Curtis;  where  the  velocity  of  the  steam  from  the  nozzles  is 
absorbed  in  and  passes  through  several  rows  of  moving  blades. 

All  other  types  are  modifications  of  these. 

There  are  certain  points  towards  which  the  effort  of  designers  should  be 
directed  in  order  to  secure  the  highest  efficiency  with  maximum  durability, 
simplicity,  and  cost  of  construction,  namely: 

First.     Reduced  steam  consumption. 

Second.     Increased  peripheral  speed. 

Third.     Simplicity  of  design. 

Fourth.     Accuracy  of  workmanship. 

Fifth.  Provision  for  expansion  and  contraction  under  all  conditions  of 
load  and  steam  pressure  in  a  manner  not  to  interfere  with  safe  operation. 

166 


STEAM   TURBINES  167 

The  small  commercial  turbines  on  the  market  to-day  are  the  de  Laval, 
Curtis,  Terry,  Kerr,  Sturtevant,  and  Dake,  all  of  the  impulse  type.  The 
old-style  de  Laval  has  only  one  row  of  moving  elements  and  one  set  of 
nozzles,  necessitating  high  bucket  speed;  but  in  their  recent  designs  there 
are  several  steam  returns. 

No  steam  turbine  should  be  thought  of  that  cannot  ultimately  meet  the 
economy  of  the  reciprocating  engine,  and  small  turbines  under  300  horse 
power  should  have  as  near  as  possible  the  economy  of  the  best  grade  of 
reciprocating  engines,  if  the  useful  field  is  to  be  extended. 

The  following  shows  the  effect  of  peripheral  speeds  on  the  economy  of 
small  turbines: 

36-inch  Curtis  19,000  feet  per  minute  —  30  pounds  per  horse  power  per  hour; 

36-inch  Curtis  25,000  feet  per  minute  —  28  pounds  per  horse  power  per  hour; 

24-inch  Terry    17,500  feet  per  minute  —  40  pounds  per  horse  power  per  hour; 

24-inch  Terry   10,000  feet  per  minute  —  50  pounds  per  horse  power  per  hour; 

24-inch  Kerr     17,500  feet  per  minute  —  40  pounds  per  horse  power  per  hour; 

24-inch  Kerr       6,000  feet  per  minute  —  60  pounds  per  horse  power  per  hour; 

30-inch  Bliss     20,000  feet  per  minute  —  40  pounds  per  horse  power  per  hour. 

Single-stage  turbines  of  the  Electra  type,  of  45  horse  power,  running  3000 
revolutions,  with  steam  from  90  to  100  pounds,  have  given  a  steam  rate  of 
from  24  to  30  pounds  per  horse  power  condensing.  This  type  is  well 
adapted  to  small  powers.  It  is  known  also  by  the  name  of  Kolb,  and  is  one 
of  the  best  for  use  in  connection  with  centrifugal  pumps.  It  is  entirely 
suitable  for  high-pressure  steam.  The  only  objection  to  it  is  the  loss  by 
friction  in  the  guide  blades,  and  the  inability  to  use  it  for  large  powers,  as 
there  is  not  enough  space  on  the  periphery  of  the  wheel  for  the  necessary 
nozzles  and  guides. 

A  mixed-flow  turbine  or  one  using  both  high-  and  low-pressure  steam  is 
one  that  will  be  valuable  in  a  great  many  places,  particularly  in  steel  mills, 
where  an  abundance  of  exhaust  steam  from  the  various  rolling-mill  engines, 
hammers,  etc.,  may  be  utilized  for  running  the  pumps.  Turbines  can  be 
built  direct-connected  with  pumps,  and  arranged  to  be  operated  either  by 
steam  at  atmospheric  pressure  and  exhaust  into  a  vacuum,  or  by  high- 
pressure  steam,  exhausting  into  a  vacuum  or  into  the  atmosphere.  Such  a 
turbine  could  be  designed  with  two  nozzle  chambers,  one  for  high-pressure 
steam  and  the  other  for  low-pressure  steam,  having  independent  governor 
control. 

The  expansion  of  steam  in  a  nozzle  obtains  a  velocity  of 


7  =  224      Hi-lHzX  +q(l-X)\, 
where  HI  =  total  head  at  PI; 

H2  =  total  head  at  P2; 
X  =  dryness  fraction; 

q  —  heat  of  liquid. 


168  CENTRIFUGAL  PUMPING  MACHINERY 

The  available  energy  in  steam  between  boiling  point  and  absolute  vacuum 
is  890,000  foot  pounds  per  pound,  and  the  velocity 

V  =  V2  X  32.2  X  890,000  =  7550  feet  per  second. 

This  is  based  upon  the  theory  advanced  as  to  molecular  velocity,  and  be- 
comes important  when  the  available  energy  in  steam  between  different 
pressures  is  to  be  determined.     The  usual  formula  for  determining  this  is  : 
Foot  pounds  per  pound  of  steam  =  778  [Hi  +  Cpti  —  (G2  +  ^2)]. 
HI  =  total  heat  at  pressure  pi\ 

Cp  =  specific  heat  at  superheated  steam  at  pressure  pi] 
ti  =  superheat  in  degrees  Fahrenheit  at  pressure  pi  ; 
Gz  =  head  of  liquid  at  pressure  p2; 
X  =  quality  of  steam  at  pressure  p2,  or  entropy; 
Vi  =  latent  heat  at  pressure  pi\ 
vz  =  latent  heat  at  pressure  p2. 
Entropy  of  superheated  steam  can  be  calculated 


of  moist  steam, 

ito  ,  . 

-Ffi  —  r  92, 
1  2 

where  T  and  TI  absolute  temperature  of  saturated  steam  at  pressures  p\ 
and  pz,  which  equals  461  (temperature  in  Fahrenheit). 
9  =  entropy  of  water  at  pressure  pi; 
92  =  entropy  of  water  at  pressure  p%. 

From  this  can  be  found  the  moisture  in  per  cent  and  available  energy  in 
foot  pounds  of  the  steam  and  the  amount  of  moisture  entering  into  the 
condensing  apparatus,  although  it  is  to  be  taken  into  account  that  some  of 
the  moisture  is  lost.  After  the  total  energy  is  found  by  taking  the  efficiency 
of  the  turbine,  the  energy  available  can  then  be  found. 

The  efficiency  of  a  turbine  is  determined  by  the  readings  of  pressure  and 
temperatures  and  the  amount  of  steam,  exhaust  pressure,  and  the  electrical 
power  of  the  generator,  and  is  the  theoretical  water  rate  divided  by  the 
observed. 

The  small  steam  turbines  now  coming  into  general  use  vary  from  10 
horse  power  to  500  horse  power.  They  are  nearly  all  of  the  impulse  type 
and  promise  to  become  the  preeminent  driving  power  for  centrifugal  or 
turbine  pumps.  The  field  for  this  combination  is  considerable  and  covers 
centrifugal  pumps,  feed  pumps,  condenser  equipments,  and  marine  auxili- 
aries. Steam  consumption  in  some  cases  is  of  importance,  but  not  in 
others. 

Steam  turbines  of  200  to  500  horse  power  may  be  obtained  having  an 
economy  equal  to  that  of  a  reciprocating  engine,  and  when  the  entire  in- 
stallation is  considered  the  steam  turbine  will  show  an  advantage. 


APPENDIX. 

ELECTRICAL  DATA. 

Full-load  speeds  for  alternating-current  motors  based  on  4  per  cent  slip. 

TABLE  NO.  8. 


< 

Cycles. 

Number  of 

poles. 

25 

27 

30 

33i 

40 

42 

50 

60 

100 

2 

1440 

1560 

1730 

1920 

2300 

2420 

2880 

4 

720 

780 

865 

960 

1150 

1210 

1440 

i725 

2880 

6 

480 

520 

575 

640 

770 

807 

960 

1150 

1920 

8 

360 

390 

433 

480 

575 

605 

720 

862 

1440 

10 

290 

310 

345 

385 

460 

485 

575 

690 

1150 

12 

240 

260 

288 

320 

385 

403 

480 

575 

960 

14 

205 

222 

247 

275 

330 

346 

412 

492 

822 

16 

180 

195 

216 

240 

287 

302 

360 

431 

720 

18 

160 

171 

192 

214 

256 

268 

320 

384 

640 

20 

142 

156 

173 

192 

230 

242 

287 

345 

575 

B.h.p.  output  for 

alternating-current  motor  = 
where 


volts  X  ampere  X  cos  <f>  X  ^n  X 


.746 
n  =  number  of  phases; 

=  power  factor  of  motor; 
m  =  motor  efficiency. 


169 


170 


CENTRIFUGAL  PUMPING  MACHINERY 


POINTS   TO   CONSIDER  IN   CENTRIFUGAL  PUMP 
INSTALLATIONS. 


Floor  Line  to  Discharge  Level  (Ft.): 

Should  be  given  in  open  system.  Pumping  overboard  or  into 
a  tank. 

Mine  Service:  Give  information  about  discharge  line  in  addi- 
tion to  above. 


Gauge  Reading  at  Discharge  Nozzle  (Lbs.): 

Should  be  given  when  pumping  into  a  closed  pressure  system. 

(Heating  System)   or  when   pumping  direct   into   city  water 
main:  Information  about  discharge  line  not  needed. 


Gauge  Reading  at  End  of  Discharge  Line  (Lbs.): 

Should  be  given  when  a  certain  pressure  is  required 
at  a  distance  away  from  the  pump ;  for  instance,  in 
sprinkler  systems  and  fire-hydrant  systems.  Infor- 
mation about  discharge  line  is  needed  in  addition  to 
above. 


Head-on  Suction: 

V///A 


Floor  Line  to  Suction  Level  (Ft.): 

Should  be  given  when  suction  water  level  is  above  floor  line. 
=1  This  is  the  case  with  all  submerged  pumps,  also  when  pump 
=•  takes  its  water  from  an  elevated  tank,  etc.  Information  about 

suction  line  is  to  be  given  in  addition  to  above. 


Gauge  Reading  at  Suction  Nozzle  (Lbs.): 

In  closed  system,  when  water  enters  under  pressure;  for  instance, 
if  connected  to  city  main,  in  heating  system,  etc.     Information 
in      about  suction  line  not  needed. 


Suction  Lift: 


Floor  Line  to  Suction  Level  (Ft.): 

This  is  the  most  common  case.  Should  always  be  given  when 
pump  takes  its  water  from  a  well  or  from  a  river,  provided  pump 
is  above  water  surface. 

Information  about  suction  pipe  needed.  State  if  foot  valve 
is  provided. 


APPENDIX 


171 


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CENTRIFUGAL  PUMPING  MACHINERY 


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APPENDIX 


173 


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174 


CENTRIFUGAL  PUMPING  MACHINERY 


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APPENDIX 


175 


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176 


CENTRIFUGAL  PUMPING  MACHINERY 


APPENDIX 


111 


Example. — To  obtain  size  of  pump  pulley  and  belt  see  belt  curve, 
page  4;  and  for  742  re  volutions^  per  minute  find  diameter  of  pulley,  assum- 


R 


per 


Min. 


\V\£3X 


\ 


\ 


X 


48 
42" 
36" 
30" 

24" 
22" 
20" 

18" 

16"       g 

14"       * 

S, 

12"       g 
1 
10" 


6" 


ing  best  belt  velocity  at  about  75  feet  per  second.     Curve  points  to 
22-inch  diameter  pulley,  belt  velocity  71  feet  per  second. 

On  the  curve  for  standard  riveted  double  belt  the  horse  power  per  inch 
of  width  for  71  feet  per  second  belt  velocity  =  10.3  horse  power.     There- 


178 


CENTRIFUGAL  PUMPING  MACHINERY 


fore  for  transmitting  84.7  horse  power  the  width  of  belt  required  is 
84.7  -r-  10.3  =  8.22  inches.  Take  a  belt  9  inches  wide  and  a  pulley  with 
a  10-inch  face. 

742  X  22 
Driving  pulley  =  —  ^n  —  =  65.3  inches,  say  66-inch  diameter  pulley, 


which  allows  for  1  per  cent  slip. 


TABLE  NO.  15.  — SPEED  OF  ROTARY  FIELD  FOR  DIFFERENT  NUMBERS 
OF  POLES  AND   FOR  VARIOUS  FREQUENCIES. 


"8 
Jg 

Speed  of  revolving  magnetism,  in  revolutions  per  minute,  when  frequency  is: 

c  — 

la 

25 

30 

33* 

40 

50 

60 

661 

80 

100 

120 

125 

133* 

2 

1500 

1870 

2000 

2400 

3000 

3600 

4000 

4800 

6000 

7200 

7500 

8000 

4 

750 

900 

1000 

1200 

1500 

1800 

2000 

2400 

3000 

3600 

3750 

4000 

6 

500 

600 

667 

800 

1000 

1200 

1333 

1600 

2000 

2400 

2500 

2667 

8 

375 

450 

500 

600 

750 

900 

1000 

1200 

1500 

1800 

1875 

2000 

10 

302 

360 

400 

480 

600 

720 

800 

960 

1200 

1440 

1500 

1600 

12 

250 

300 

333 

400 

500 

600 

667 

800 

1000 

1200 

1250 

1333 

14 

214 

257 

286 

343 

428 

514 

571 

686 

857 

1029 

1071 

1143 

16 

188 

225 

250 

300 

375 

450 

500 

600 

750 

900 

938 

1000 

18 

167 

200 

222 

267 

333 

400 

444 

533 

667 

800 

833 

889 

20 

150 

180 

200 

240 

300 

360 

400 

480 

600 

720 

750 

800 

22 

136 

164 

182 

217 

273 

327 

364 

436 

545 

655 

682 

720 

24 

125 

150 

167 

200 

250 

300 

333 

400 

500 

600 

625 

667 

TABLE  NO.   16.  — SLIP  OF  INDUCTION  MOTORS. 


Slip,  at  full  load,  per  cent. 

Slip,  at  full  load,  per  cent 

Capacity  of 
motor,  H.P. 

Capacity  of 
motor,  H.P. 

Usual  limits. 

Average. 

Usual  limits. 

Average. 

i 

20-40 

30 

15 

5-11 

8 

1 

10-30 

20 

20 

4-10 

7 

1 

10-20 

15 

30 

3-9 

6 

I 

8-20 

14 

50 

2-8 

5 

2 

8-18 

13 

75 

1-7 

4 

3 

8-16 

12 

100 

1-6 

3.5 

5 

7-15 

11 

150 

1-5 

3 

7? 

6-14 

10 

200 

1-4 

2.5 

10 

6-12 

9 

300 

1-3 

2 

APPENDIX 


179 


FRICTION  OF  WATER  IN  PIPES. 

The  following  tables  show  the  flow  of  water  in  pipes  1000  feet  long, 
as  computed  by  the  Hazen-  Williams  formula  V=  1.32  Cr63  S-64, 
Where  V  =  velocity  of  water  in  feet  per  second. 
C  =  an  empirical  coefficient. 

r  =  the  hydraulic  radius  in  feet  =  ^  for  pipes  running  full. 

,,     i    j      i.     i  h       head 

the  hydraulic  slope 


S 


The  friction  heads  are  given  for  C 


, 
100  only. 


For  other  values  of  C  multiply  by  A  = 


TABLE  NO.  17.  — FRICTION  HEAD  IN  FEET  PER  1000  FEET  LENGTH 

OF  PIPE. 

XOTE.  —  Figures  in  this  Table  give  the  friction  head  in  feet  per  1000  feet  length  of  pipe  based  on  factor  C 
100  for  pipes  of  condition  or  age  represented  opposite  factor  C  =  100  in  Table  3. 


|  Approximate.  1 

C  =  100 

Size  of  pipe  in  inches. 

Standard  wrought-iron  pipe. 

2J 

3 

4 

5 

| 

i  j  t. 

i 

i 

1 

H 

14 

2 

Inside  d\a.  d  =      .27 

.364 

.494 

.623 

.824 

1.048 

1.38 

1.611 

At  const,  bead 
cap.  is   prop.     .0378 
tod*-*  =           | 

.08 

.172 

.306 

.616 

1.12 

2.24 

3.3 

5.7 

9.9 

15.6 

32 

56 

At  const,  capac- 

^    hrooV3  700,000 
prop,  to  -j-  = 

157,000 

34,000 

10,700 

2,630 

790 

200 

92.5 

31.25 

10.2 

4.12 

.977 

32 

|  U.  8.  gallonsTper  minute  =  O. 

.2 
.5 
1 
1.5 
2 
3 
4 
5 
6 
8 
10 
15 
20 
30 
40 
50 
60 
80 
100 
150 
200 
300 
400 
500 
600 

62 
345 
1210 

78 
280 
600 
1030 

64 
138 
233 
490 
840 
1260 

21 
45 
74 
158 
270 
410 
570 
980 
1470 

19 
41 
70 
105 
147 
250 
380 
800 
1360 

12.6 
21.4 
32.5 
45.5 
78 
117 
250 
420 
890 
1520 

For  other  values  of  C 
multiply  figures  in  Table  by  A. 

5.7 
8.4 
12 
20.3 
30.5 
65 
111 
235 
400 
600 
850 
1450 

2.62 
3.98 
5.6 
9.5 
14.3 
30 
52 
110 
188 
284 
396 
680 
1020 

2 
3.3 
5 
10.7 
18.2 
38.4 
66 
99 
139 
237 
358 
760 
1290 

1.1 
1.7 
3.6 
6.1 
12.9 
22 
33.2 
46.5 
79 
120 
256 
431 
920 
1560 

.7 
1.5 
2.5 
5.4 
9.1 
13.8 
19.2 
32.8 
49.6 
105 
178 
380 
650 
980 

.62 
1.32 
2.23 
3.39 
4.72 
8.1 
12.2 
26 
44 
93 
160 
240 
337 

.44 
.75 
1.13 
1.59 
2.71 
4.11 
8.7 
14.8 
31.4 
54 
81 
113 

At  constant  diameter  head  ia  approximately  proportional  to  G*, 


180 


CENTRIFUGAL  PUMPING   MACHINERY 


TABLE  NO.  18.  —  TABLE  GIVING  COEFFICIENT  OF  CONDITION  C, 

FACTOR  a  AND  CORRESPONDING  YEARS  OF  SERVICE  IN  SOFT, 

CLEAR,  UNFILTERED  RIVER  WATER. 


C 

r/ioo, 

V  c 

A 

Diameter  of  pipe  in  inches. 

i-H 

2-3 

4 

5 

8 

8 

10 

12 

16 

20 

24 

30 

36 

42 

48 

54 

60 

140 
130 

120 
110 
100 

90 
80 
60 
40 

Condition  of  pipe. 

Years  of  service  of  cast-iron  pipe. 

.54 
.62 

.71 
.84 
1 
1.22 
1.52 
2.58 
5.46 

Very  smooth 
and  straight  W.  I. 

Smooth  new  wroti 

Do.  brass,  tin,  etc. 

Ordinary  straight 
brass,  tin,  etc. 
ght  iron 

00 

0 
4 

00 

0 
4 

00 

0 
4 

00 

0 
5 
10 
16 

00 

0 
5 

10 
17 

00 

0 
5 
10 
17 

26 
37 

00 

0 
5 

11 
18 

27 
3D 

00 

0 
5 

11 
19 

28 
41 

00 

0 

5 
11 
19 

29 
42 

00 

0 

6 
12 
19 

30 
43 

00 

0 
6 

12 
20 

30 
44 

00 

0 

6 
12 
20 

30 
45 

00 

0 

6 
12 
20 

31 
46 

00 

0 
6 
12 
20 
31 
47 

Ordinary  Wrougl 

Old  wrought  iron 
Very  rough 

it  Iron 

13 

14 

15 

Very  rough 
Badly  tubercu- 
lated 

26 
45 

75 

28 
50 

87 

30 
55 

95 

33 
62 

35 
08 

00  means  Very  Best  and  New  Cast  Iron  Straight  Pipe.        0  means  Good  New  Cast  Iron  Pipe. 

TABLE  NO.   19.— FRICTION  HEAD   IN   FEET   PER   1000   FEET  LENGTH 

OF  PIPE. 

NOTE.  —Figures  in  this  Table  give  the  friction  head  in  feet  per  1000  feet  length  of  pipe  based  on  factor  C  = 
100  for  pipes  of  age  represented  opposite  factor  C  =  100  in  Table  3. 


« 
< 

C  =  100 

Size  of  pipe  in  inches. 

6 

8 

10 

12 

16 

20 

24 

30 

36 

42 

48 

54 

60 
27,900 

At  const,  head 
cap.   is   prop, 
to  d2-5  = 

88.2 

181 

316 

499 

1024 

1790 

2820 

4930 

7780 

11,430 

16,000 

21,400 

A1 

it? 
1 

const,  capac- 
r  head  is  prop. 
100  Mill 

12,900 

3,050 

1,000 

402 

954 

31.25 

12.6 

4.12 

1.65 

.765 

.403 

.218 

.129 

# 

3 
g 

a 

1 

13 

3 

QQ 

D 

35 
42 
56 
70 
105 
140 
210 
280 
350 
420 
560 
700 
1,400 
2,100 
2.800 
3,500 
4,200 
5.600 
7,000 
10,500 
14,000 
21,000 
28,000 
35,000 
42,000 
56,000 
70,000 

.24 
.33 
.57 
.86 
1.84 
3.1 
6.6 
11.3 
16.9 
23.8 
40.4 
61 
222 

.13 
.22 
.43 
.77 
1.62 
2.76 
4.18 
5.9 
9.9 
15.1 
54 
116 

.07 
.16 
.26 
.55 
.93 
1.41 
1.97 
3.38 
5.1 
18.4 
38.6 
66 

.03 
.07 
.11 
.22 
.38 
.58 
.81 
1.38 
2.1 
7.6 
16.0 
27 
41 
58 
99 
150 

.026 
.056 
.095 
.143 
.201 
.34 
.52 
1.87 
3.98 
6.8 
10.2 
14.3 
24.2 
36.8 
78 

.032 
.049 
.068 
.115 
.174 
.63 
1.33 
2.28 
3.43 
4.81 
8.2 
12.4 
26.2 
44.8 

For  other  values  of  C 
multiply  figures  in  Table  by  A. 

.02 
.03 
.05 
.072 
.259 
.55 
.93 
1.41 
1.97 
3.38 
5.1 
10.8 
18.3 
39 

.024 
.087 
.184 
.315 
.476 
.67 
1.13 
1.72 
3.64 
6.2 
13.2 
22.4 

.036 
.076 
.129 
.196 
.274 
.467 
.71 
1.49 
2.55 
5.4 
9.2 
13.9 
19.6 

.036 
.061 
.092 
.129 
.22 
.332 
.7 
1.21 
2.56 
4.35 
6.6 
9.2 
15.7 

.032 
.048 
.068 
.115 
.174 
.37 
.63 
1.33 
2.28 
3.44 
4.8 
8.2 
12.4 

.018 
.028 
.038 
.065 
.098 
.21 
.354 
.75 
1.28 
1.94 
2.71 
4.61 
7 

.011 
.016 
.023 
.039 
.059 
.125 
.212 
.449 
.76 
1.16 
1.62 
2.78 
4.19 

At  constant  diameter  head  is  approximately  proportional  to  G3  and  capacity  is  approximately  proportional 


to 


APPENDIX  181 

APPROXIMATE  RULES  FOR  INTERPOLATING. 

RULE  1.  —  At  constant  head,  capacity  is  proportional  to  d?-b. 
Example:    A  4-inch  pipe  discharges  60  gallons,  how  much  would  a 
1-inch  pipe  discharge  under  the  same  conditions? 

60  X  1  12 
42-5  =  32,  1.048"  =1.12,  hence  capacity  =  -  -  =  2.1  g.p.m. 

tM 

RULE  2.  —  At  constant  capacity,  head  is  proportional  to  -=• 

Example:  Capacity  600  g.p.m.,  diameter  =  3  inches.     What  is  the  head? 
The  nearest  figure  in  Table  is  337  feet  for  a  4-inch  pipe,  and  the  friction 

337  X  4.12 
in  a  3-inch  pipe  must  be  greater,  hence  h  =  =  1420  feet. 


RULE  3.  —  At  constant  diameter,  head  is  proportional  to  G2. 

Example:  Same  as  before. 

The  nearest  figure  in  table  is  980  feet  head  for  500  g.p.m.,  and  the 

980  X  6002 
friction  for  600  g.p.m.  must  be  greater,  hence  h  =  --      2        =  1410  feet. 

RULE  4.  —  At  constant  diameter,  capacity  is  proportional  to  vh. 
Example:  Diameter  20  inches,  h  =  64  feet  per  1000  feet  length.     What 
is  the  capacity? 

The  nearest  figure  in  Table  is  44.8  feet  giving  14,000  g.p.m.     The  capacity 

14,000  X  A/64      14,000X8 

will  be  greater,  hence  -  -  =  -  *-=-=  --  =  16,700  g.p.m. 

V44.8  6.7 


INDEX. 


A. 

Angles,  determining,  60. 

Application  of  analysis  to  problem,  52. 

theory,  41. 

theory  ta  problem,  69. 
for  low-lift  pumps,  2. 
Appold's  formula,  46. 
Automatic  repriming.  10. 
Automobile  fire  engines,  140. 

B. 

Bearings,  lubrication  of,  28. 

Belfast  sewage  pump,  98. 

Biro  tor-type  pumps,  145. 

Blades,  46. 

Boiler  feeding,  107. 

Brooklyn  high-pressure  plant,  134. 

C. 

Capacity,  46. 
Capacity-head  curves,  19. 
Centrifugal  dredgers,  101. 

pumping  unit  at  Montreal,  87. 
Centrifugal- jet  condensers,  112. 
Characteristics,  short  method  of  finding, 

58. 

Chart ,  graphical,  64. 
Circulation  of  water,  109. 
Circumferential  velocity,  68. 
Condenser  plants,  109. 
Condenser,  centrifugal-jet,  112. 
Correcting  impeller- vane  angle,  65. 
Curtis  turbine,  166. 

D. 

Deep  mining  pumps,  104. 
Definition  of  slip,  74. 
De  Laval  turbine,  166. 
Determining  angles,  60. 

turbine  efficiency,  168. 
width  of  impeller  at  circum- 
ference, 68. 
Diffusion  guides  and  vanes,  theory  of,  39. 

vanes,  5,  159. 
Direct-connected  engines,  165. 


Discharge  piping,  27. 
Drainage  work,  90. 
Dredging  pumps,  100. 
Dry-dock  pumps,  116. 
Dry  irrigating  pumps,  96. 

E. 

Efficiency  tests,  22. 
Egypt,  pumping  plants  in,  90. 
Electric  motors,  161. 
Engine-driven  pumps,  164. 
Erection,  27. 
Excavating  pumps,  101. 

F. 

Fireboats,  142. 
Fire  engines,  140. 
pumps,  132. 

Fire-station  service,  129. 
Foot  valves,  12,  28. 
Foundations,  27. 

Francis  formulae  for  weir  calculation,  148. 
Friction  losses,  15,  53. 

G. 

Gas-engine-driven  pumps,  139. 
Graphical  chart,  64. 

H. 

High-lift  pumps,  2. 

High-pressure  fire-service  plant  in  Brook- 
lyn, 134. 

High-speed  pumps,  148. 
Horizontal  turbine  pump,  157. 
Hot-well  pumps,  110. 
Hydraulic  mining  pumps,  100. 
Hydraulic  thrusts,  7. 

I. 

Impeller-vane  angle,  method  of  correct- 
ing, 65. 

Impellers,  theory  of,  32,  48,  66. 
Installation,  27. 

of  waterworks,  81 .       « 
Irrigating  pumps  in  India,  96. 


183 


184 


INDEX 


J. 

Jennings  Canal  Co.  plant,  91. 

L. 

League  Island  plant,  127. 
Losses  by  friction,  15,  53. 
Losses  in  efficiency,  16. 
Low-lift  pumps,  1. 
Lubrication  of  bearings,  28. 

//    M. 
Method  of  co/recting  impeller-vane  angle, 

65. 

Methods  of  obtaining  total  head,  14. 
Mining  pumps,  104. 
Montreal  waterworks,  tests  at,  86. 
Motors,  electric,  161. 
Movable  diffusors,  6. 
Multistage  pumps,  4,  158. 

N. 
Norfolk  dry-dock,  120. 

P. 

Panama  Canal  dredging  plant,  101. 

Paper  mill  pump,  155. 

Parsons  turbine,  166. 

Piping,  27. 

Plant  of  Jennings  Canal  Co.,  91. 

Pontoon  dock  pumps,  117. 

Priming,  28. 

devices,  8. 

Propeller  pumps,  73. 
Pumps,  dredging,  100. 

for  boiler  feeding,  107. 

for  dry-docks,  116. 

for  irrigation,  96. 

high-lift,  2. 

high-speed,  148. 

low-lift,  1. 

multistage,  4. 

on  shipboard,  144. 

propeller,  73. 

turbine,  83. 

screw,  73. 

turbine  underwriter  fire,  129. 

volute  type,  1,  12. 


R. 


Relief  valves,  12. 


S. 

Screw  pumps,  73. 

Sewerage  pumping  plant  in  Belfast,  98. 
Shipboard  pumps,  144. 
Short  method  of  finding  characteristics,  58. 
Skin  friction,  16. 
Slip,  74. 

Speed  regulation,  163. 
Starting,  28. 
Strainers,  28. 
Stuffing  boxes,  28. 
Suction  foot  valve,  28. 

piping,  27. 
Sugar  house  pump,  154. 

T. 
Tees  dock  pumps,  Middlesbrough,  Eng., 

127. 
Tests  at  Montreal  waterworks,  86. 

for  series,  20. 
Theory  of  centrifugal  pump,  31. 

diffusion  guides  and  vanes,  39. 
impellers,  32,  48,  66. 
Thrust  bearings,  158. 
Total  head,  methods  of  obtaining,  14. 
Trirotor  pump,  150. 
Turbine,  principal  types,  166. 
pumps  for  mines,  106. 
underwriter  fire  pumps,  129. 

V. 

Vanes,  theory  of,  39. 
Valves,  foot,  12. 

relief,  12. 

Velocity,  circumferential,  68. 
Vertical  house  sump  pump,  157. 

turbine  pump,  156. 
Volute  type  pumps,  1,  12. 

W. 

Waterworks  installation,  81. 

Whirlpool  chambers,  5. 

Width  of  impeller  at   circumference,   to 

determine,  68. 
Worthington  centrifugal  pump,  91. 

Z. 

Zoelly  &  Rateau  turbine,  166. 


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